Power transmission elements, torque measuring device and freewheel assembly

ABSTRACT

A measuring device with a crankshaft and a load cell for determining a radial force acting on the crankshaft having a receiving sleeve for receiving a bearing ring and a fastening ring for attaching the load cell in a transmission housing. Axial support areas are provided on the fastening ring for axially supporting the outer ring of the first bearing. Moreover, measuring regions for receiving radial forces of the receiving sleeve are provided which connect the receiving sleeve with the fastening ring. Strain sensors are attached to at least two of the measuring regions. An evaluation electronics is connected to the strain sensors.

CROSS-REFERENCE TO RELATED APPLICATIONS

The present application is a continuation of U.S. application Ser. No.16/462,744, filed May 21, 2019, which is hereby specificallyincorporated by reference herein in its entirety.

TECHNICAL FIELD

The present description discloses in a first aspect a load cell fordetermining a radial force acting on a crankshaft.

BACKGROUND

US 2007/051187 discloses a load sensing bearing assembly which comprisesa flange arrangement with spring regions which serve for determiningradial forces, thrust forces and tilting moments on an outer ring of abearing.

U.S. Pat. No. 4,900,165 A discloses a bearing support system whichreceives an axial thrust load of a bearing of a shaft in a controlledmanner by retaining the bearing with a slotted inner ring which issupported with a radially spaced outer ring by means of a plurality ofcircumferentially spaced supports which are radially stiff and axiallyflexible.

US 2011/187179 A1 discloses a wheel bearing unit, having a first elementwhich comprises a cylindrical section which is supported by a bearingarrangement and which comprises a radially extending flange section andhaving a second element which comprises flaps for centering a wheel rim.

SUMMARY

The present description discloses in a first aspect a load cell fordetermining a radial force acting on a crankshaft. The load cellcomprises a cylindrical receiving sleeve for receiving a ring of abearing, and a fastening ring for attaching the load cell in atransmission housing. The ring may in particular be an outer ring of arolling bearing. The fastening ring of the load cell is connected to thereceiving ring via connecting areas or measuring regions. The measuringregions are provided for receiving radial forces of the receivingsleeve, which are transmitted from the ring of the bearing to themeasuring regions.

Strain sensors are mounted in at least two of the measuring regions, forexample as glued strain gauges. Furthermore, the load cell comprisesaxial support areas, which are provided on the fastening ring for axialsupport of the outer ring of the bearing or for receiving axial forces.Here, axial forces are aligned parallel to the longitudinal direction ofthe receiving sleeve and radial forces perpendicular thereto.

In particular, the load cell may be configured to receive a bearingwhich is located radially inside the receiving sleeve, wherein an outerring of the bearing is in contact with an inner surface of the receivingsleeve and the measuring regions and the axial support areas may joinradially inwardly to the fastening ring.

The axial support areas are separated from the measuring regions byradial slots, wherein the axial support areas are separated from thereceiving sleeve by a circumferential slot.

In particular, the measuring regions may comprise measuring lugs formedas angle brackets, wherein the angle brackets may in particular comprisea radial region and an axial region adjoining the radial region. Thisshape is well suited for producing a controlled deformation under theaction of radial forces.

The radial region is connected to the fastening ring and the axialregion is connected to the receiving sleeve. In particular, the radialregion may be arranged to the axial region at an angle of approximately90 degrees.

According to a further embodiment, the axial region is flush with acylindrical inner surface of the receiving sleeve.

In particular, the axial support lugs may be configured such that theyproject radially inwardly over an inner surface of the receiving sleeve.As a result, a bearing can be introduced into the receiving sleeve suchthat a ring of the bearing, in particular an outer ring, abuts againstthe projecting region and the axial forces acting on the bearing aretransferred to the load cell and transferred via the fastening region.

According to a further embodiment, at least one of the strain sensors isconfigured as a strain gauge. According to a further embodiment, astrain sensor is attached to each of the measuring regions, inparticular the strain sensor may each be attached to a radial region ofthe measuring region.

According to a further embodiment, at least two of the measuring regionscomprise lowered areas for attaching the strain sensors, whereby thestrain sensors can be easily positioned and protected from damage.

According to a special embodiment, the load cell comprises fourmeasuring regions, which are arranged at intervals of 90 degrees. As aresult, on the one hand, a good support of the bearing received in theload cell can be achieved and, on the other hand, a measurement ofdefined radial forces can be made possible, from which a torque actingon a crankshaft can be calculated, the crankshaft being supported by thebearing.

In particular, the fastening ring may comprise fastening regions inwhich fixing holes are provided, which are suitable for screwing thefastening ring to a housing.

Furthermore, the fastening ring may comprise recesses, wherein themeasuring regions are arranged in the recesses. In this case, themeasuring lugs, the slots that delimit the measuring lugs and the axialsupport areas may particularly be dimensioned such that an angular rangewhich comprises the measuring lugs and the slots delimiting themeasuring lugs corresponds approximately to an angular range of theaxial support areas.

According to an embodiment, an angular extent of the measuring regionsin a circumferential direction is less than or equal to 30 degrees. As aresult, a good deformability can be achieved, and the applied radialforce corresponds to a well-defined direction.

In particular, the load cell may be made integrally of metal, whichelastically deforms and provides good stability and stability of shape.Here, “metal” refers to a metallic material and includes, in particular,metal alloys.

In a further aspect, the present description discloses a measuringdevice for determining a force acting on a crankshaft. The measuringdevice comprises a crankshaft with a bearing arranged on the crankshaftand with the load cell described above. The receiving sleeve of the loadcell is arranged on an outer ring of the bearing, and the axial supportareas of the load cell are supported in an axial direction on the outerring of the bearing.

Furthermore, an evaluation electronics is connected to the strainsensors of the load cell, which is configured to determine a forceacting on the crankshaft from signals of the strain sensors. Thecrankshaft may in particular be a pedal shaft, but for example also acrankshaft of a piston engine.

Furthermore, the present description discloses a transmissionarrangement with the aforementioned measuring device. The transmissionarrangement comprises a transmission housing and a crankshaft, whereinthe crankshaft is mounted in the transmission housing via a firstbearing and a second bearing.

The first bearing is received in the transmission housing via the loadcell of the measuring device. The load cell is in turn received in thetransmission housing via the fastening ring. The receiving sleevereceives an outer ring of the first bearing and the axial support areasare supported on the outer ring of the first bearing.

In the transmission arrangement, in particular the crankshaft maycomprise a first step and a second step for supporting the bearings. Thestep scan be formed, for example, in that the crankshaft has a largerdiameter in the middle than at its two ends. In particular, an innerring of the first bearing of the measuring device may rest against thefirst step of the pedal shaft and an inner ring of the second bearingmay rest against the second step of the crankshaft such that anX-arrangement of an obliquely mounted bearing is formed.

In the X-arrangement of the bearings arranged on the crankshaft forcesacting on the crankshaft are transferred outwards via the bearings. As aresult, a load cell can be arranged further outside in the vicinity ofan evaluation electronics. In addition, a force acting on an outputshaft arranged on the output side bearing also acts outwards. Thereby,the output shaft can be kept at a distance from a rotor shaft, which isarranged at a drive side of the output shaft.

For receiving axial forces, the first bearing of the measuring deviceand the second bearing of the measuring device may be configuredspecifically as single-row angular contact ball bearings. According toan embodiment, the second bearing is supported by a wave spring on thesecond step of the pedal shaft or on the housing.

According to a further embodiment, the second bearing is supported by aspacer disk on the second step of the pedal shaft or on the housing.

Furthermore, the present description discloses a transmissionarrangement having the above features, further comprising a motor and areduction gear connected to the motor, and a hollow output shaftconnected to the reduction gear.

In this transmission arrangement, the crankshaft is configured as pedalshaft, and the first bearing and the second bearing are each configuredas rolling bearings. The pedal shaft is passed through the hollow outputshaft, and a freewheel is provided between the pedal shaft and thehollow output shaft for decoupling the pedal shaft from the hollowoutput shaft.

Furthermore, the present description discloses an electrically drivenvehicle with the aforementioned transmission arrangement. Here, themotor is configured as an electric motor, and a battery of theelectrically driven vehicle is connected to the electric motor.

In another aspect, the present description discloses a harmonic pin ringtransmission with an input shaft and an output shaft. In particular, theinput shaft may be configured as a hollow shaft, which is suitable as arotor shaft of an electric motor, and the output shaft may be configuredas a hollow shaft, which is located in the flow of forces after thebelow-mentioned inner gear.

In this transmission, a cam disk which serves to press a tractionmechanism to outer gears, is formed in one piece with a hollow driveshaft, wherein the hollow drive shaft may in particular form a rotor ofan electric motor.

The transmission comprises a first outer gear and an inner gearconcentrically disposed with respect to the first outer gear in a firstaxial plane. A second outer gear is disposed in a second axial plane anda traction mechanism extends between the first outer gear and the innergear. In other words, an axial region of the traction mechanism islocated in a space formed between the outer gear and the inner gear.

A revolving transmitter lifts the traction mechanism from an outercircumference of the inner gear and presses it against an innercircumference of the first outer gear and against an inner circumferenceof the second outer gear. The transmitter comprises a hollow drive shaftand a cam disc. The cam disk and a pin retaining ring of the pin ringare arranged in a third axial plane, which is located between the firstaxial plane and the second axial plane. The cam disk is formed in onepiece with the hollow drive shaft.

According to an advantageous embodiment, the traction mechanism isconfigured as a pin ring, wherein pins or projections protrude on twoopposite sides of a central region, wherein the central region isarranged in the third axial plane. In particular, the central region maybe made flexible and corresponds to a pin retaining ring. Furthermore,the central region of the pin ring may comprise an inner bearing surfacefor bearing on a cam disk and an outer bearing surface for bearing on agear part, such as a support ring.

In operation, the revolving transmitter lifts the pins from an outercircumference of the inner gear and presses the pins against an innercircumference of the first outer gear and against an inner circumferenceof the second outer gear.

According to another embodiment, a circumference of the cam disk isconfigured to have an oval shape, such as a circular shape superimposedby a sine wave or an ellipse. According to another embodiment, acircumference of the cam disk has a circular shape and is arrangedeccentrically to a transmission central axis.

In particular, a rolling bearing may be arranged radially between thecam disk and the traction mechanism to avoid forces along thecircumference, wherein the rolling bearing is advantageously deformablefor an oval cam disk.

In particular, the transmitter may essentially consist of lightweightaluminum and may be manufactured in one piece with a hollow shaft whichforms a rotor of an electric motor.

In particular, if the transmitter comprises an eccentric disc or aneccentrically arranged circular disc, the transmitter may comprise aring which is connected with the hollow drive shaft via connectingstruts.

In another aspect, the present description discloses a harmonic pin ringtransmission having a support ring.

The transmission comprises the following components. A first outer gear,an inner gear disposed concentrically with respect to the first outergear in a first axial plane, and a second outer gear disposed in asecond axial plane.

Further, the transmission comprises a traction mechanism extendingbetween the first outer gear and the inner gear, and a revolvingtransmitter which lifts the traction mechanism from an outercircumference of the inner gear, presses it against an innercircumference of the first outer gear and against an inner circumferenceof the second outer gear.

The transmitter comprises a hollow drive shaft and a cam disk, whereinthe cam disk and a pin retaining ring of the pin ring are arranged in athird axial plane, which is located between the first axial plane andthe second axial plane. Further, the first outer gear is formed by afirst outer ring, and the second outer gear is formed by a second outerring, wherein the first outer ring and the second outer ring areinserted into a support ring. In particular, the second outer ring maybe identical in construction to the first outer ring and/or may bemirror symmetrical.

Furthermore, the first outer ring and the second outer ring may each bemade of plastic such as PA66 or polymethyl methacrylate (PMMA) byinjection molding. It is particularly possible to form the outer gearsof plastic, when the pins of a pin ring abut all teeth of the outer geartoothing, so that there is a uniform distribution of the load. By asuitable dimensioning of the pins of a pin ring and a suitably selectedtoothing each of the pins can abut either on the internal toothing ofthe outer gear or on the external toothing of the inner gear.

In particular, the first outer ring and the second outer ring may eachcomprise radially outwardly projecting journals which are distributedover the circumference of the respective outer ring, wherein the supportring comprises matching recesses into which the journals are inserted.

According to another embodiment, grooves are formed in the first outerring and the second outer ring, which are respectively distributed overthe circumference of the first outer ring and the second outer ring, andthe support ring is formed as a region of a transmission housingcomprising journals which engage in the grooves. This embodiment isparticularly suitable for a geared motor.

The support ring which receives the outer gears may be made oflightweight aluminum, in particular it may be made by an aluminumdie-casting process.

According to a further embodiment, the support ring comprises twosubrings, which abut each other in an axial direction. In this way, thehousing parts mutually support each other in the axial direction.

For easier connecting to the transmission housing, the first outer ring,the second outer ring and the support ring may have matching and alignedscrew holes. In particular, the outer gears or outer rings and thesupport ring may be connected to the transmission housing by screwsbeing passed through screw holes of a transmission cover and through thematching screw holes of the first outer ring, the support ring and thesecond outer ring and screwed into a thread of a transmission housing ofthe harmonic pin ring transmission.

In another aspect, the present description discloses a harmonictransmission with a traction mechanism, and more particularly a harmonicpin ring transmission with a pin ring comprising a freewheel device.

The transmission comprises an input shaft for applying a driving forceof a motor and an output shaft for transmitting an output force.Furthermore, the transmission comprises a first outer gear, an innergear, which is arranged concentrically with respect to the first outergear in a first axial plane, and a second outer gear, which is arrangedin a second axial plane, and a traction mechanism, which extends betweenthe first outer gear and the inner gear.

Furthermore, the transmission comprises a revolving transmitter whichlifts the traction mechanism from an outer circumference of the innergear, presses it against an inner circumference of the first outer gearand presses it against an inner circumference of the second outer gear.

The transmitter comprises a hollow drive shaft and a cam disk, whereinthe cam disk and the traction mechanism are arranged in a third axialplane, which is located between the first axial plane and the secondaxial plane. When using a pin ring, a pin retaining ring of the pin ringor a portion of a one-piece pin ring, which corresponds to a pinretaining ring, is arranged in the third axial plane.

Furthermore, the transmission optionally comprises a hollow outputshaft, which is supported in the inner gear via a motor freewheel, and apedal shaft, which is supported in the hollow output shaft via a pedalshaft freewheel. The pedal shaft is received in a transmission housingof the transmission. The pedal shaft comprises a receiving region or aninterface for the motor freewheel on an outer circumference. At an innercircumference opposite to the outer circumference, the pedal shaftcomprises a receiving region for the pedal shaft freewheel.

In particular, the motor freewheel may be configured as a clamp rollerfreewheel and the pedal shaft freewheel may be configured as a pawlfreewheel. According to a further exemplary embodiment, both the motorfreewheel and the pedal shaft freewheel are configured as clamp rollerfreewheels.

The output shaft or hollow output shaft extends in the axial directionon the output side of the hollow drive shaft, wherein a ball bearing isarranged between the hollow output shaft and the pedal shaft, andwherein the hollow output shaft comprises a fastening region for anoutput member such as a gear or pulley.

Furthermore, the present description discloses a freewheel assemblyhaving an outer clamp roller freewheel and an inner pawl freewheel.

The freewheel assembly comprises a hollow drive shaft, a hollow outputshaft and a pedal shaft. In this arrangement, the pedal shaft, thehollow output shaft and the hollow drive shaft are arrangedconcentrically with each other. The hollow output shaft is arrangedradially inside the hollow drive shaft and the pedal shaft is arrangedradially inside the hollow output shaft.

Further, the hollow output shaft comprises a stair-shaped pawlengagement region on an inner circumference and comprises a stair-shapedclamp roller rolling region on an outer circumference which is radiallyopposite to an inner circumference.

The pedal shaft comprises a star-shaped receiving region for pawls,wherein the star-shaped receiving region comprises pawl seats beingevenly distributed over the circumference for receiving pawls and springseats being arranged adjacent to the pawl seats for receiving springs.

In this arrangement, steps of the stair-shaped pawl engagement regionand steps of the stair-shaped clamp roller rolling region formed on thepedal shaft are aligned such that a drive direction of the outer clamproller freewheel coincides with a drive direction of the inner pawlfreewheel.

The outer clamp roller freewheel comprises, inter alia, the hollow driveshaft and the stair-shaped clamp roller rolling region of the hollowoutput shaft and the inner pawl freewheel comprises, inter alia, thepedal shaft and the stair-shaped pawl engagement region.

In the present description, “radially inward” with respect to a hollowshaft refers to the inner circumference or its imaginary extension. Inthis case, the component does not need to be completely within thehollow shaft in the axial direction.

According to a further embodiment, the stair-shaped clamp body rollingregion on the outer circumference of the hollow output shaft and thestair-shaped pawl engagement region on the inner circumference of thehollow output shaft are located essentially in the same axial plane. Asa result, a tilting moment on the hollow output shaft can be avoided andspace can be saved in the axial direction.

According to a further embodiment, the hollow drive shaft of thefreewheel assembly comprises a disk-shaped region with an externaltoothing, which is provided on an outer circumference of the disk-shapedregion. The disk-shaped region need not be formed as a solid disk, butmay for example be formed as a disk which is perforated or comprisesother interruptions or as a ring with a strut. The disk-shaped regionserves to receive an output torque. For example, it can be configured asan inner gear of a harmonic pin ring transmission.

According to a further embodiment, the hollow output shaft comprises anannular thickening at a first end and comprises a fastening region foran output means, in particular for a chainring adapter, at a second endopposite the first end.

According to a further embodiment, the outer circumference of the hollowoutput shaft of the freewheel assembly comprises a steplike bearingregion for a rolling bearing. Accordingly, the inner circumference ofthe hollow output shaft may also comprise a steplike bearing region fora rolling bearing.

According to a further embodiment, the hollow output shaft comprises aninternal thread on an output side end.

According to a further embodiment, the freewheel assembly comprisespawls which are rotatably mounted in the pawl seats and spring elementswhich are arranged in the spring seats and are in contact with thepawls.

Furthermore, the freewheel assembly comprises a freewheel cage with websand clamp rollers arranged between the webs, wherein the freewheel cageand the clamp rollers are arranged radially between the clamp rollerrolling region of the hollow output shaft and an inner circumference ofthe hollow drive shaft.

According to a further embodiment, the pawl seats are cylindricallyshaped, closed at one end by a wall and are open at an opposite end.

According to a further embodiment, the stair-shaped clamp body rollingregion and the freewheel cage each comprise at least two receivingregions for spring elements, such as helical springs, wherein in eachcase a spring element is arranged between a receiving region of theclamp body rolling region and a receiving region of the freewheelingcage.

Furthermore, the pedal shaft may comprise a force sensor unit, whereinthe force sensor unit comprises a metal load cell and a drive-side pedalshaft ball bearing and wherein the load cell is arranged on the pedalshaft ball bearing.

According to a further embodiment, the load cell of the freewheelassembly comprises an inner annular portion which is fastened on anouter annular portion via fastening lugs, which may be in particularfour fastening lugs arranged at a 45° spacing. The pedal shaft ballbearing is inserted into the inner annular portion.

According to a further embodiment, the inner portion and the outerportion of the load cell are offset radially to each other, wherein thefastening lugs are laterally bounded by radial slots, and wherein atleast two of the fastening lugs comprise a strain sensor. The fasteninglugs are suitable for receiving radial forces and are also referred toas measuring lugs.

Furthermore, an axial thickness of the outer ring may be reduced in theregion of the fastening lugs, so that the fastening lugs are set backrelative to an annular fastening region which serves for fastening to atransmission housing.

Furthermore, the present application discloses a pedal shaft for afreewheel assembly, the pedal shaft comprising a first fastening regionfor a pedal crank at a first end and a second fastening region for apedal crank at a second end opposite thereto. Furthermore, the pedalshaft comprises a star-shaped receiving region for pawls in the vicinityof the first end.

According to a particular embodiment, the star-shaped receiving regioncomprises steps, the steps each comprising a first side surface, asecond side surface, a pawl support region inclined in a predefineddirection relative to the circumferential direction by approximately45°, and a spring seat. Furthermore, the steps comprise an upper surfacethat is substantially parallel to the circumference of the shaft.

In addition, the steps comprise an end region with a pawl seat, whereinthe pawl seat is at least partially cylindrical, axially open to oneside and closed to an axially opposite side. In particular, there may besix steps.

In another aspect, the present description discloses a harmonictransmission, in particular a harmonic pin ring transmission, having anoutput shaft which comprises integrated interfaces for freewheels.

The harmonic transmission comprises an input shaft and an output shaftand further comprises the following features.

A first outer gear and an inner gear disposed concentrically withrespect to the first outer gear in a first axial plane. A second outergear is arranged in a second axial plane. Furthermore, a tractionmechanism is provided, which extends between the first outer gear andthe inner gear, for example a pin ring.

A revolving transmitter is connected to the input shaft. In operation,the transmitter lifts the traction mechanism from an outer circumferenceof the inner gear and presses it against an inner circumference of thefirst outer gear and against an inner circumference of the second outergear.

The transmitter comprises a hollow drive shaft and a cam disk, whereinthe cam disk and, if the traction mechanism is configured as a pin ring,a pin retaining ring of the pin ring are arranged in a third axial planelocated between the first axial plane and the second axial plane.

Furthermore, the harmonic transmission comprises a hollow output shaft,which is supported in the inner gear via a motor freewheel, and a pedalshaft, which is supported in the hollow output shaft via a pedal shaftfreewheel. The pedal shaft is received in a transmission housing of theharmonic transmission. Further, the pedal shaft comprises a receivingregion or an interface for the motor freewheel on an outer circumferenceand a receiving region for the pedal shaft freewheel on an innercircumference radially opposite to the outer circumference.

According to an embodiment, the engine freewheel is configured as aclamp roller freewheel and the pedal shaft freewheel is configured as apawl freewheel.

The output shaft extends in axial direction on the output side of thehollow drive shaft, wherein a ball bearing is arranged between thehollow output shaft and the pedal shaft, and wherein the hollow outputshaft comprises a fastening region for an output element, in particularan output element such as a gear or pulley for connection to a tractionmechanism.

In another aspect, the present description discloses a harmonic pin ringtransmission having an input shaft and an output shaft, the transmissionhaving the following components.

In the transmission, a first outer gear and an inner gear are provided,wherein the inner gear is arranged concentrically with respect to thefirst outer gear in a first axial plane. A second outer gear is arrangedin a second axial plane. A pin ring with pins extends between the firstouter gear and the inner gear.

A revolving transmitter is arranged in the area of the inner gear andthe outer gears. In operation, the revolving transmitter lifts the pinsof the pin ring from an outer circumference of the inner gear andpresses the pins against an inner circumference of the first outer gearand against an inner circumference of the second outer gear.

The transmitter comprises a hollow drive shaft and a cam disk, whereinthe cam disk and a central region of the pin ring are arranged in athird axial plane, which is located between the first axial plane andthe second axial plane.

The pin ring is made of one piece. In particular, the pin ring may bemade of metal. The pins of the pin ring will be formed by projectionsprotruding from two axially opposite sides of a central region of thepin ring, the central region comprising a smooth, cylindrical innerbearing surface for bearing on a bearing and a smooth, cylindrical outersurface.

According to a further embodiment, the projections are cylindrical on afirst side of the two opposite sides and the projections are partiallycylindrical on a second of the two opposite sides, wherein acylindrically formed region is located in the radial direction on theoutside of the pin ring.

According to a further embodiment, the projections comprise on a firstside of the two opposite sides an inner rounded engagement region in theradial direction on the inside of the pin ring and an outer roundedengagement region in the radial direction on the outside of the pinring. Furthermore, the projections comprise on a second of the twoopposite sides an outer rounded engagement region.

According to a further embodiment, a bearing such as a rolling or ballbearing or a flexible rolling or ball bearing is arranged between thecam disk and the pin ring, and wherein the pin ring comprises on aninner side a shoulder for supporting the bearing.

In another aspect, the present description discloses a one-piece pinring for a harmonic pin ring transmission which is formed in one pieceand made of metal. The pin ring comprises pins that are formed byprojections that protrude from two axially opposite sides of a centralregion of the pin ring.

In this case, the central region comprises a smooth, cylindrical innerbearing surface for bearing on a bearing and a smooth, cylindrical outerbearing surface for bearing on a support cylinder.

According to a further embodiment, the projections are cylindrical on afirst side of the two opposite sides and the projections are partiallycylindrical on a second of the two opposite sides. A cylindricallyformed region is located in the radial direction on the outside of thepin ring.

According to a further embodiment, the projections comprise on a firstside of the two opposite sides an inner rounded engagement region in theradial direction on the inside of the pin ring and an outer roundedengagement region in the radial direction on the outside of the pinring. Furthermore, the projections comprise on a second of the twoopposite sides an outer rounded engagement region.

According to a further embodiment, a tangentially arranged web islocated between the projections in each case on the first of theopposite two sides, wherein an outer boundary line of a cross section ofthe projections smoothly merges into an outer boundary line of the web.Herein, “smooth” can be understood to refer to a continuous firstderivative when averaging over surface roughness.

According to a further embodiment, a bearing such as a rolling or ballbearing, or a flexible rolling or ball bearing is arranged between thecam disk and the pin ring. In this case, the pin ring comprises on aninner side a shoulder for supporting the bearing.

In another aspect, the present description discloses a harmonictransmission with an obliquely mounted bearing.

The transmission comprises a first outer gear and an inner gearconcentrically disposed with respect to the first outer gear in a firstaxial plane and a second outer gear disposed in a second axial plane.

A traction mechanism extends between the first outer gear and the innergear. Furthermore, a revolving transmitter is provided, which lifts thetraction mechanism from an outer circumference of the inner gear andpresses it against an inner circumference of the first outer gear andagainst an inner circumference of the second outer gear.

The transmitter comprises a hollow drive shaft and a cam disk, whereinthe cam disk and a pin retaining ring of the pin ring are arranged in athird axial plane, which is located between the first axial plane andthe second axial plane.

A pedal shaft is arranged radially inside the output shaft, wherein thepedal shaft is mounted in the motor housing via a drive-side pedal shaftball bearing and a load cell consisting of metal or a metal alloy.

According to a further embodiment, the load cell comprises an innerannular portion which is attached to an outer annular portion viafastening lugs, wherein the pedal shaft ball bearing is inserted intothe inner annular portion. In particular, these may be four fasteninglugs spaced 45 degrees apart. The fastening lugs are also referred to asmeasuring lugs.

The outer annular portion is inserted into a cylindrical portion of themotor housing. The inner portion and the outer portion are radiallyoffset from each other. Furthermore, the fastening lugs are boundedlaterally by radial slots and a material recess is provided radiallyoutside of the measuring lugs. As a result, among other things, thethickness of the measuring lugs is reduced, whereby a deformability ofthe measuring lugs is facilitated.

On a surface of at least two of the measuring straps, strain sensors,such as strain gauges, may be applied on at least two of the measuringstraps.

Furthermore, a wave spring may be arranged between the load cell and thedrive-side rotor ball bearing, which can serve, inter alia, thecompensation of tolerances.

In another aspect, the present description discloses a harmonic pin ringtransmission with a crank gear formed as a planetary gear. The harmonicpin ring transmission comprises an input shaft and an output shaft,which are also referred to as drive shaft and driven shaft.

Furthermore, the transmission comprises a first outer gear and an innergear, which is arranged concentrically with respect to the first outergear in a first axial plane, and a second outer gear, which is arrangedin a second axial plane.

A traction mechanism extends between the first outer gear and the innergear. A revolving transmitter is provided to lift the traction mechanismfrom an outer circumference of the inner gear and to press it against aninner circumference of the first outer gear. The transmitter comprises ahollow drive shaft and a cam disk, wherein the cam disk is arranged in athird axial plane located between the first axial plane and the secondaxial plane.

Furthermore, a pedal shaft is arranged radially inside the output shaft,and a planetary gear and a pedal shaft freewheel are arranged in theflow of forces between the pedal shaft and the output shaft.

According to an embodiment, a planet carrier of the planetary gear isconnected to the pedal shaft, a ring gear of the planetary gearcomprises a connection region for connection to a transmission housingand a sun gear of the planetary gear is mounted on the pedal shaft. Thepedal shaft freewheel is disposed between a hollow shaft of theplanetary gear which is connected to the sun gear and the output shaft.

According to another embodiment, the pedal shaft freewheel is arrangedbetween the crankshaft and a planet carrier of the planetary gear,wherein a ring gear of the planetary gear is rotatably mounted in theharmonic transmission and wherein a sun gear of the planetary gear isadapted for attachment to a stationary housing part of the harmonictransmission.

Furthermore, the present description discloses a tension shafttransmission, wherein the tooth geometry can be designed in particularso that there is a complete tooth engagement between the inner and outerteeth. This is the case even if the transmitter comprises an eccentricdisk instead of an oval cam disk.

The tension shaft transmission has two main types, namely: first, adesign with driven cup-shaped tensioning shaft, which comprises afastening region for an output shaft and second, a design with movablesecond outer gear and cylindrical tensioning shaft.

The tension shaft transmission comprises an outer gear with internaltoothing, the outer gear comprising a fastening region for attachment toa transmission housing, and a tensioning shaft having an externaltoothing, the tensioning shaft being concentrically disposed withrespect to the outer gear in an axial plane.

A revolving transmitter is configured to be suitable for attachment to arotating input shaft and to be capable of, during operation, pressingthe tensioning shaft against the internal toothing of the outer gear.For this purpose, the transmitter is disposed within the tensioningshaft and an outer circumference of the transmitter is suitable forpressing the tensioning shaft.

The transmitter comprises a hollow drive shaft and a cam disk, whichpreferably comprises a circular circumference arranged eccentrically tothe central axis of the outer gear or an oval or ellipticalcircumference arranged centered to the central axis of the outer wheel,wherein a ball bearing is arranged on the circumference of the cam diskwhich in the case of the oval or elliptical circumference is preferablya flexible ball bearing. However, the toothing of the outer gearspecified below is particularly suitable for an oval transmitter andcauses in this case a particularly good meshing.

A cross section of the tooth crests of the external toothing of thetensioning shaft corresponds essentially to a sector of a circle. Thus,the cross section of the tooth crests corresponds to a cross section ofa portion of cylindrical pins and preferably essentially to asemicircle.

With respect to a central axis of the outer gear, the internal toothingof the outer gear is essentially defined by an outer equidistant to thegear trajectory that is defined by the formulasx(t)=r1*cost(t)+r2*cos((n+1)*t)+r3*cos((n+3)*t) andy(t)=r1*sin(t)r2*sin((n+1)*t)+r3*sin((n+3)*t), wherein n+1 is the numberof teeth of the internal toothing of the outer gear, wherein the radiir1, r2 and r3 are greater than zero, and wherein for the scale of theradii r2>r3 and r1>r2+r3 applies. Further conditions for the parametersn and t and for the radii are given below in the description inconnection with this toothing.

The coordinates x and y relate to a Cartesian coordinate system arrangedperpendicular to the central axis of the outer gear with its origin inthe central axis of the outer gear or the transmission.

In particular, the tensioning shaft may be formed in a cup shape,wherein at the bottom of the cup shape, a fastening region is formed forfastening an output shaft. This can also be done in such a way that thetensioning shaft is formed integrally with the output shaft.

Furthermore, a central circular opening may be formed at the bottom ofthe cup shape, wherein the fastening region of the tensioning shaftcomprises fastening holes arranged around the central circular opening.This embodiment may be advantageous in particular for a geared motor.

According to an alternative embodiment, the tensioning shaft has theshape of a circular cylinder, wherein the transmission comprises asecond rotatably arranged outer gear which comprises a fastening regionfor fastening an output shaft, wherein the internal toothing of theouter gear is determined by the same construction or formula as theinternal toothing of the first outer gear.

The expression “essentially”, in particular with respect to a toothing,may for example refer to max. 5% or 10% deviation with respect to thedistances mentioned in the description with reference to FIGS. 117, 118.

Alternatively, the shape of the outer gear toothing can also bespecified by explicitly specifying the epicyclic construction, accordingto which the tooth surface of the internal toothing of the outer gear isdetermined by a radial distance from a central axis of the inner gear asa function of a cycle angle. In this case, the radial distance from thecentral axis in turn is determined by an inner equidistant to a geartrajectory, wherein a location on the gear trajectory is determined ineach case by the vector sum of a cycle vector, a first epicycle vectorand a second epicycle vector.

Further, a tail of the cycle vector is located on the central axis, atail of the first epicycle vector in the tip of the cycle vector, and atail of the second epicycle vector in the tip of the first epicyclevector.

Furthermore, an epicycle angle of the first epicycle vector is n+1 timesthe cycle angle and an epicycle angle of the second epicycle vector n+3times the cycle angle, where n is a number of pins of the harmonic pinring transmission which is at least four.

Furthermore, a length of the cycle vector is greater than the sum of thelengths of the first epicycle vector and the second epicycle vector, anda length of the first epicycle vector is greater than a length of thesecond epicycle vector.

Furthermore, the present description discloses a two-stage reductiongear comprising an outer gear fixedly secured to the transmissionhousing with a first internal toothing, wherein the outer gear comprisesa fastening region for attachment to a transmission housing.

An outer gear rotatably mounted on the transmission housing is providedwith a second internal toothing, wherein the outer gear comprises afastening region for attachment to an output shaft.

A two-part or two-section, but nevertheless one-piece pin ring isarranged concentrically to the outer gears, wherein the two-partone-piece pin ring comprises a first external toothing and a secondexternal toothing. The first external toothing of the two-part one-piecepin ring engages in the internal toothing of the stationary outer gear.The second external toothing of the two-part one-piece pin ring engagesin the internal toothing of the rotatable outer gear.

A revolving transmitter is configured to press the two-part one-piecepin ring against the internal toothing of the stationary outer gear andagainst the internal toothing of the rotatable outer gear. A ballbearing is arranged on a circumference of the revolving transmitter.

In particular, to achieve a high reduction, the number of teeth of theinternal toothing of the stationary outer gear may be greater than thenumber of teeth of the first external toothing, and the number of teethof the internal toothing of the rotatable outer gear may be greater thanthe number of teeth of the second external toothing, wherein further,the number of teeth of the stationary outer gear is greater than thenumber of teeth of the rotatable outer gear and the number of teeth ofthe first external toothing is greater than the number of teeth of thesecond external toothing.

Furthermore, the transmitter may comprise a circular ring arrangedeccentric to the axis of the stationary outer gear. The eccentrictransmission offers a specifically simple and robust design that doesnot require a deformable bearing.

In particular, a cross section of the tooth crests of the first externaltoothing and a cross section of the tooth crests of the second externaltoothing may essentially correspond to a sector of a circle, preferablya semicircle. Thus, the toothing geometry corresponds to cylindricalpins, which is particularly favorable in the eccentric transmission.

Furthermore, a cross section of the tooth crests of the first externaltoothing and a cross section of the tooth crests of the second externaltoothing may essentially correspond to an inner equidistant, inparticular at a distance of a pin radius, a gear trajectory defined bythe formula x(t)=r1*cos(t)+r2*cos(nt) and y(t)=r1*sin(t)+r2*sin(nt),wherein the following applies for the radii r1, r2: r1>0, r2>0 andr1>r2. In particular, the toothing may correspond to this form along theentire circumference, and not only in the region of the tooth crests.

The same applies to the toothing of the pin ring. Thus, a cross sectionof the tooth crests of the first external toothing and a cross sectionof the tooth crests of the second external toothing may essentiallycorrespond to an outer equidistant, in particular at the distance of apin radius, to the gear trajectory that is defined by the formulax(t)=r1*cos(t)+r2*cos(nt) and y(t)=r1*sin(t)−r2*sin(nt), wherein thefollowing applies for the radii r1, r2: r1>0, r2>0 and r1>r2, and wherer1, r2 and n have the same values as for the outer gear toothing. Inparticular, this toothing form may apply to the toothing along theentire circumference, and not only in the region of the tooth crests.

Furthermore, the present description discloses a load cell fordetermining a radial force acting on a crankshaft with a receivingsleeve for receiving a ring of a bearing, a fastening ring for attachingthe load cell in a transmission housing, and axial support areasprovided on the fastening ring for axially supporting the ring of thebearing.

Furthermore, measuring regions for receiving radial forces of thereceiving sleeve are provided which connect the receiving sleeve withthe fastening ring, wherein strain sensors are attached to at least twoof the measuring regions.

Furthermore, the present description discloses a freewheel assemblyhaving an outer transmission freewheel and an inner pedal shaftfreewheel. The freewheel assembly comprises a hollow drive shaft, ahollow output shaft, and a pedal shaft. The pedal shaft, the hollowoutput shaft and the hollow drive shaft are arranged concentrically witheach other.

Furthermore, the hollow output shaft is disposed radially inside thehollow drive shaft, and the pedal shaft is disposed radially inside thehollow output shaft, wherein the pedal shaft freewheel is arrangedbetween the pedal shaft and the hollow output shaft. The transmissionfreewheel is arranged opposite the pedal shaft freewheel on the hollowoutput shaft. The hollow output shaft comprises adapted areas on aninner side and on an outer side in the region of the respectivefreewheel.

Thus, the double freewheel can be integrated in the drive in aspace-saving manner in the region of the pedal shaft without the needfor a separate outer ring or inner ring. In this case, the outerfreewheel may in particular be a clamp roller freewheel and the innerfreewheel may particularly be a pawl freewheel. However, both freewheelsmay also be clamp roller freewheels. Other combinations are alsopossible.

Furthermore, the present description discloses a one-piece pin ring,which is preferably used in conjunction with an eccentric transmission.

According to a first embodiment, the one-piece pin ring is made ofmetal, wherein a pin retaining ring and an arrangement of pins, whichprotrude in axial direction on two opposite sides of the pin retainingring, are made in one piece.

In particular, the pins may be connected to each other in thecircumferential direction, which provides additional stability and canenable a more efficient production.

Furthermore, the pins may be formed on a first of the two opposite sidesas half pins, which are suitable for engagement in an internal toothing,and the pins on a second of the opposite sides may be formed as wholepins, which are suitable for engagement in an internal toothing and forengagement in an external toothing opposite the internal toothing. Thus,weight and material can be saved.

In a further embodiment, the one-piece pin retaining ring comprises asmooth inner circumference on an inner side and comprises rounded bulgeson an outer side, which are made in one piece with the pin retainingring. This embodiment is suitable for example for a two-stage reductiongear with eccentric.

Furthermore, at least one head region of the rounded bulges may comprisea semicircular cross section. Thus, the same toothings can be used,which are also suitable for a pin ring with cylindrical pins.

Furthermore, the present description discloses a support ring assemblyfor a reduction gear having a first outer gear and a second outer gearcomprising a support ring, a first outer gear having a first internaltoothing and a second outer gear having a second internal toothing,wherein the first outer gear and the second outer gear are inserted intothe support ring on opposite sides, and wherein the support ringcomprises a fastening region, for example axial bores, for attachment toa transmission housing.

In particular, the first outer gear and the second outer gear may bemade of plastic. Furthermore, the first outer gear and the second outergear may each be connected with the support wheel via a pin-grooveconnection, so that they can be easily assembled.

Furthermore, the present description discloses a one-piecerotor-transmitter element for a reduction gear comprising a hollow shaftcomprising a fastening region on a first side for fastening a rotorpackage, and comprising a cam disk on a second side opposite to thefirst side, wherein an outer circumference of the cam disk is configuredas a receiving area for a ball bearing.

In particular, the one-piece rotor-transmitter element may be made ofaluminum. In addition, the hollow shaft of the one-piecerotor-transmitter element may be dimensioned such that a pedal shaft canbe passed through the hollow shaft.

In a further embodiment, the cam disk comprises a circular circumferencearranged eccentrically relative to the central axis of the hollow shaft.In an alternative embodiment, the cam disk comprises an ovalcircumference relative to the central axis of the hollow shaft.

Furthermore, the present description discloses crank gears that providea speed-up transmission. According to a first embodiment, the crank gearcomprises a drive shaft, in particular a crankshaft or pedal shaft witha planetary gear arranged on the drive shaft, wherein a planet carrierof the planetary gear is fixedly connected to the drive shaft, a ringgear of the planetary gear comprises a fastening region for attachmentto a transmission housing and a receiving region for a torque sensor,and a sun gear of the planetary gear is configured as a ring gear, whichis arranged concentrically to the drive shaft, and wherein the sun gearis connected to a hollow output shaft of the planetary gear, which isrotatably mounted on the drive shaft.

According to a second embodiment, the crank gear comprises a driveshaft, in particular a crankshaft or pedal shaft, with a planetary geararranged on the drive shaft, wherein a planet carrier of the planetarygear is mounted on the drive shaft via a freewheel, a sun gear of theplanetary gear comprises a fastening region for attachment to atransmission housing and a receiving region for a torque sensor, and aring gear of the planetary gear comprises a receiving region for a ballbearing for supporting on a transmission housing.

According to a third embodiment, the crank gear comprises a drive shaft,in particular a crankshaft or pedal shaft, with a planetary geararranged on the drive shaft, wherein a planet carrier of the planetarygear comprises a fastening region for attachment to a transmissionhousing, wherein a hollow shaft of the planetary gear is fixedlyconnected to the drive shaft is, and wherein a sun gear of the planetarygear is configured as a hollow shaft which is arranged concentrically tothe drive shaft and rotatably mounted on the drive shaft.

Furthermore, the present description discloses a cycloidal gear, thecycloidal gear comprising the following components: a transmissionhousing, an outer gear having an internal toothing which is fixed to thetransmission housing, and an input shaft arranged concentric with theouter gear, wherein the input shaft comprises a drive-side eccentricdisk, on which a first ball bearing is arranged, and an output-sideeccentric disk, on which a second ball bearing is arranged.

A drive-side inner gear with an external toothing is mounted on thefirst ball bearing and an output-side inner gear with an externaltoothing is mounted on the second ball bearing. The drive-side innergear and the output-side inner gear are disposed inside the outer gear,and the external toothings of the drive-side inner gear and theoutput-side outer gear respectively engage with the internal toothing ofthe outer gear.

In particular, the cycloidal gear may comprise a crankshaft, which ismounted within the input shaft, and the load cell described above, whichis mounted on the crankshaft on the drive side.

Furthermore, the input shaft may be configured as a one-piece rotortransmitter element described above.

According to another embodiment, the cycloidal gear comprises acrankshaft mounted within the input shaft, and wherein the crankshaftcomprises one of the planetary gears described above, wherein thecrankshaft forms the drive shaft of the planetary gear.

According to a further embodiment, a third ball bearing is disposed onthe input shaft on the output side of the output side eccentric disk,wherein a driven pulley is arranged on the third ball bearing, whereinthe driven pulley comprises carrier pins which engage in axiallysuccessively arranged openings of the drive side inner gear and theoutput side inner gear, wherein an output shaft is formed radiallyinside on the driven pulley.

In this case, an output shaft is formed radially inside on the drivenpulley, wherein the third ball bearing is arranged on an inner shoulderof the output shaft. Furthermore, an inner gear ball bearing is arrangeddiagonally opposite with respect to the centers of the bearings in theaxial direction to the third ball bearing on an outer shoulder of theoutput shaft, wherein the inner gear ball bearing is supported on ahousing cover of the transmission housing.

According to a further alternative embodiment, at least one of the innergears comprises a first toothing and a second external toothing. Thecycloidal gear further comprises a rotatable outer gear having aninternal toothing, wherein the second external toothing engages in theinternal toothing of the rotatable outer gear, and wherein the rotatableouter gear comprises a fastening region for mounting an output shaft.Thus, a two-stage reduction can be provided.

Further, the internal toothing may be formed by an inner surface of theouter gear or formed by an arrangement of stationary pins on whichrollers are arranged.

Disclosed is a load cell for determining a radial force acting on acrankshaft, the load cell comprising: a receiving sleeve for receiving aring of a bearing; a fastening ring for attaching the load cell in atransmission housing; axial support areas provided on the fastening ringfor axially supporting the ring of the bearing; and measuring regionsfor receiving radial forces of the receiving sleeve and which connectthe receiving sleeve with the fastening ring, wherein strain sensors areattached to at least two of the measuring regions.

Also disclosed is a transmission arrangement comprising a measuringdevice for determining a force acting on a crankshaft, the measuringdevice comprising: a crankshaft with a bearing arranged on thecrankshaft; a load cell comprising: a receiving sleeve for receiving aring of a bearing; a fastening ring for attaching the load cell in atransmission housing; axial support areas provided on the fastening ringfor axially supporting the ring of the bearing; and measuring regionsfor receiving radial forces of the receiving sleeve and which connectthe receiving sleeve with the fastening ring, wherein strain sensors areattached to at least two of the measuring regions, wherein the receivingsleeve of the load cell is arranged on an outer ring of the bearing, andwherein the axial support areas of the load cell are supported in anaxial direction on the outer ring of the bearing; and an evaluationelectronics which is connected to the strain sensors of the load cell.

Additionally, disclosed is a measuring device for determining a forceacting on a crankshaft, the measuring device comprising a crankshaftwith a bearing arranged on the crankshaft, a load cell for determining aradial force acting on a crankshaft, the load cell comprising: areceiving sleeve for receiving a ring of the bearing, a fastening ringfor attaching the load cell in a transmission housing, axial supportareas provided on the fastening ring for axially supporting the ring ofthe bearing, measuring regions for receiving radial forces of thereceiving sleeve and which connect the receiving sleeve with thefastening ring, wherein strain sensors are attached to at least two ofthe measuring regions, wherein the receiving sleeve of the load cell isarranged on an outer ring of the bearing, and wherein the axial supportareas of the load cell are supported in an axial direction on the outerring of the bearing, an evaluation electronics which is connected to thestrain sensors of the load cell.

BRIEF DESCRIPTION OF THE DRAWINGS

The subject matter of the description will be explained in more detailbelow with reference to the figures below.

Wherein

FIG. 1 is a cross-sectional view of a harmonic pin ring transmission,

FIG. 2 shows a drive-side part of an exploded view of the transmissionof FIG. 1,

FIG. 3 shows an output-side part of an exploded view of the transmissionof FIG. 1,

FIG. 4 is an exploded view of an inner portion of the transmission ofFIG. 1,

FIG. 5 shows a perspective view of a transmission assembly of thetransmission of FIG. 1,

FIG. 6 shows a bottom bracket bearing assembly of the transmission ofFIG. 1,

FIG. 7 is a cross-sectional view of a bottom bracket bearing assemblyshown in FIG. 6,

FIG. 8 shows a drive-side perspective view of the transmission of FIG. 1in the assembled state,

FIG. 9 shows an output-side perspective view of the transmission of FIG.1 in the assembled state,

FIG. 10 shows a transmission of axial forces into the housing accordingto the mounting concept of the transmission of FIGS. 1 to 9,

FIG. 11 a cross-sectional view of a harmonic pin ring transmission withan eccentric disk,

FIG. 12 is an exploded view of the transmission of FIG. 11,

FIG. 13 is a side view of a subassembly of the transmission of FIG. 11,

FIG. 14 is a cross-sectional view of the subassembly of FIG. 13,

FIG. 15 shows an output side plan view of a pin ring of the transmissionof FIG. 11,

FIG. 16 shows a cross section through the pin ring of FIG. 15,

FIG. 17 shows a partial view of the pin ring of FIG. 15,

FIG. 18 shows an output-side plan view of the pin ring of FIG. 15, and

FIG. 19 shows an output-side plan view of the transmission of FIG. 11,

FIG. 20 shows a torque curve in the transmission of FIG. 11,

FIG. 21 shows a perspective view, seen from the drive side, of a loadcell,

FIG. 22 is a plan view, seen from the drive side, of the load cell,

FIG. 24 is a cross-sectional view of the load cell along thecross-sectional line A-A of FIG. 22,

FIG. 25 shows a side view of the load cell,

FIG. 26 is a perspective view of the load cell seen from the outputside,

FIG. 27 shows a plan view of the load cell seen from the output side,

FIG. 28 is a cross-sectional view along intersection line A-A of FIG.27,

FIG. 29 is a cross-sectional view along intersection line B-B of FIG.27,

FIG. 30 shows a harmonic pin ring transmission with the load cell ofFIG. 21,

FIG. 31 shows a section of the pin ring transmission of FIG. 30,

FIG. 32 shows a cross-sectional view of an obliquely mounted bearingwith a load cell,

FIG. 33 shows a first cross-sectional view of an obliquely mountedbearing with a load cell,

FIG. 34 shows a further cross-sectional view of an obliquely mountedbearing with a load cell,

FIG. 35 shows a first embodiment of a torque sensor,

FIG. 36 shows a second embodiment of a torque sensor,

FIG. 37 shows a third embodiment of a torque sensor,

FIG. 38 shows a fourth embodiment of a torque sensor,

FIG. 39 shows a fifth embodiment of a torque sensor with two pairs ofopposing strain gauges at a 90 degree interval,

FIG. 40 shows a fifth embodiment of a torque sensor with two adjacentpairs of opposing strain gauges,

FIG. 41 shows a sequence of measurement values of a force measuringdevice according to FIG. 38,

FIG. 42 shows a further sequence of measurement values of a forcemeasuring device according to FIG. 37 compared to a direct torquemeasurement at a deformable cell,

FIG. 43 is a cross-sectional view of another HPD transmission with aload cell similar to the HPD transmission of FIG. 29,

FIG. 44 shows a detail enlargement of the HPD transmission shown in inFIG. 43 in the area of the load cell,

FIG. 45 is an exploded view of a freewheel assembly,

FIG. 46 shows an output-side view of the freewheel assembly of FIG. 44with the side cover removed,

FIG. 47 is a cross-sectional view of the free wheel assembly of FIG. 45,pedal shaft and transmission inner gear,

FIG. 48 shows an output-side perspective view of the freewheel assemblyof FIG. 45 with the side cover removed,

FIG. 49 shows a drive-side perspective view of the freewheel assembly ofFIG. 45 with the side cover removed,

FIG. 50 is a side view of the pedal crank of the freewheel assembly ofFIG. 45,

FIG. 51 is a side view of the freewheel assembly FIG. 45 with outputshaft,

FIG. 52 shows a top view of the pedal shaft in the axial direction,

FIG. 53 shows an output-side top view of the output shaft,

FIG. 54 shows a drive-side top view of the output shaft,

FIG. 55 shows a freewheel cage of the transmission freewheel,

FIG. 56 shows a clamping body ring of the transmission freewheel,

FIG. 57 shows a side view of the freewheel assembly with output shaftand sensor arrangement,

FIG. 58 shows a further cross-sectional view of the freewheel assemblyalong the cross-sectional line D-D of FIG. 57,

FIG. 59 shows a cross-sectional view, seen from the output side, of thefreewheel assembly of FIG. 57,

FIG. 60 is a cross-section of a harmonic pin ring transmission with aclamp roller freewheel,

FIG. 61 is a cross-sectional view of a geared motor with a harmonic pinring transmission taken along the intersection line of FIG. 61,

FIG. 62 shows a side view of the geared motor on the output side,

FIG. 63 shows a drive-side part of an exploded view of the geared motorof FIG. 60,

FIG. 64 shows an output-side part of the exploded view of the gearedmotor of FIG. 60,

FIG. 65 shows a perspective view of the geared motor of FIG. 60 seenfrom the output side, and

FIG. 66 shows a perspective view of the geared motor of FIG. 60 seenfrom the output side,

FIG. 67 shows a harmonic pin ring transmission with a planetary geararranged on a pedal shaft, in which an output takes place via a planetcarrier,

FIG. 68 shows a perspective view of the planetary gear assembly of FIG.67,

FIG. 69 shows a side view of the planetary gear assembly of FIG. 68,

FIG. 70 shows a cross section along the intersection line C-C of FIG. 69

FIG. 71 shows a cross section along the intersection line B-B of FIG.69,

FIG. 72 shows a harmonic pin ring transmission with a planetary geararranged on a pedal shaft, in which an output takes place via a ringgear,

FIG. 74 shows a side view of the planetary gear assembly of FIG. 68,

FIG. 75 shows a cross section along the intersection line C-C of FIG.74, and

FIG. 76 shows a cross section along the intersection line B-B of FIG.74,

FIG. 77 is a cross-sectional view of a motor gear unit for an electricbicycle with cycloidal gear,

FIG. 78 is an exploded view of the motor gear unit of FIG. 77,

FIG. 79 is a side view of the cycloidal gear of FIG. 77 on the outputside,

FIG. 80 shows a cross section of the cycloidal gear shown in FIG. 77along the intersection line A-A of FIG. 79,

FIG. 81 is the side view of FIG. 79, in which components hidden in FIG.79 are indicated by dashed lines,

FIG. 82 shows a cross section of a motor gear unit with a tension shafttransmission,

FIG. 83 shows an exploded view of the tension shaft transmission of FIG.82,

FIG. 84 shows a drive-side side view of the tension shaft transmissionof FIG. 82,

FIG. 85 shows a cross section along the cross-sectional line A-A of FIG.84,

FIG. 86 shows a cross section along the cross-sectional line B-B of FIG.84,

FIG. 87 shows a detail enlargement of the detail “C” of FIG. 84,

FIG. 88 shows a drive-side perspective view of the tension shafttransmission in the assembled state,

FIG. 89 shows a side view of a two-stage transmission gearing with atwo-part pin ring,

FIG. 90 is a cross-sectional view of the two-stage reduction gearingtaken along the cross-sectional line A-A of FIG. 89,

FIG. 91 is the side view of FIG. 89, in which components hidden in theview of FIG. 89 are indicated by dashed lines,

FIG. 92 is an output side of a three-quarter section of the two-stagereduction gearing of FIG. 82,

FIG. 93 shows an output-side perspective view of the three-quartersection of FIG. 92,

FIG. 94 shows an HPD-F transmission with an oval cam disk,

FIG. 95 shows an HPD-E transmission with a single eccentric disk,

FIG. 96 shows an epicyclic construction for an inner gear toothing of anHPD-E transmission,

FIG. 97 shows an epicyclic construction for an outer gear toothing of anHPD-E transmission,

FIG. 98 shows an epicyclic construction for an inner gear toothing of anHPD-F transmission,

FIG. 99 shows an epicyclic construction for an outer gear toothing of anHPD-F transmission,

FIG. 100 shows an application of the epicyclic construction of FIG. 98in a first angular position,

FIG. 101 shows the epicyclic construction of FIG. 100 in a secondangular position,

FIG. 102 shows the epicyclic construction of FIG. 100 in a third angularposition,

FIG. 103 shows the epicyclic construction of FIG. 100 in a fourthangular position,

FIG. 104 shows an application of the epicyclic construction of FIG. 97in a first angular position,

FIG. 105 shows the epicyclic construction of FIG. 11 in a second angularposition,

FIG. 106 shows the epicyclic construction of FIG. 11 in a third angularposition,

FIG. 107 shows the epicyclic construction of FIG. 11 in a fourth angularposition,

FIG. 108 shows a tooth geometry for the transmission of FIG. 2 accordingto the construction of FIG. 3 and FIG. 4,

FIG. 109 is a top view of the transmission of FIG. 95 on the outputside,

FIG. 110 shows a cross-sectional view of the transmission of FIG. 2,

FIG. 111 is a top view of the transmission of FIG. 95 on the drive side,

FIG. 112 shows a toothing geometry for another HPD-E transmission with55 pins and the associated gear trajectories according to theconstruction of FIG. 96 and

FIG. 97,

FIG. 113 shows a toothing geometry for a HPD-F transmission with 150pins according to the construction of FIG. 98 and FIG. 99,

FIG. 114 shows a toothing geometry of the inner gear toothing of FIG.113 and the associated gear trajectory,

FIG. 115 shows a worn out toothing of a transmission according to FIG.94 and an accordingly adapted tooth geometry,

FIG. 116 shows a detail of an HPD-F transmission with a toothingaccording to the epicyclic constructions shown in FIGS. 98 and 99 and anintegrally formed pin ring,

FIG. 117 shows a tolerance range of a tooth profile defined byenvelopes, and

FIG. 118 shows a tolerance range of a tooth profile defined by profiledisplacements,

FIG. 119 shows a schematic drawing of a first drive,

FIG. 120 shows a schematic drawing of a further drive,

FIG. 121 shows a schematic drawing of a strain wave drive

FIG. 122 shows a schematic drawing of a harmonic pin ring drive,

FIG. 123 shows a schematic drawing of an eccentric drive,

FIG. 124 shows a driven-side view of a motor gear unit with a strainwave gearing which is similar to the strain wave gearing of FIG. 83,

FIG. 125 is a perspective cross-sectional view of the motor gear unit ofFIG. 124, cut along intersection line A-A of FIG. 124, and

FIG. 126 shows a cross section along the intersection line A-A of FIG.124.

DETAILED DESCRIPTION

In the following description, details are provided to illustrate theembodiments of the specification. However, it should be apparent tothose skilled in the art that the embodiments can be implemented withoutsuch details.

FIG. 1 shows a cross-sectional view of a harmonic pin ring gear 10. Thecross-sectional plane A-A of FIG. 1 is marked in FIG. 9. In FIG. 1, theleft side corresponds to a drive side and the right side corresponds toan output side of the harmonic pin ring gear 10. According to the usualarrangement of the drive on the right side, the viewing direction ofFIG. 10 is directed in the direction of travel.

A stator 20 of a stator assembly of the harmonic pin ring gear 10 isdisposed in a motor housing 22. The stator 20 comprises three separatecoils 21 for connection to three phases of a three-phase inverter. Thethree coils of the stator 20 are connected to the three-phase invertervia three terminals 25, one of which is shown in FIG. 1.

The three-phase inverter is configured as a power electronics, which isarranged on a printed circuit board 23, wherein the printed circuitboard 23 is arranged in a cooling cover 24, which is mounted on themotor housing 22 on the drive side. The printed circuit board 23 isconfigured as an annular disk, which is located outside of cylindricalprotruding portions of the motor housing and the cooling cover, so thatthe motor electronics located thereon is sealed against oil and greaseof the transmission.

A pedal shaft 35 extending centrally through the motor housing 22 isstepped on the output side and comprises three steps whose diameterincreases from outside to inside and on each of which the shaft seal 50,the output side pedal shaft ball bearing 46 and the pedal shaftfreewheel 49 are arranged. The diameter of the pedal shaft 35 is alsostepped on the drive side and comprises two steps, wherein on the outerstep, the shaft seal and a sensor ring 68 are arranged, and on the innerstep, the ball bearing 45 is arranged. The stair-shape of the pedalshaft is also shown in the perspective view of FIG. 4.

An outer rotor shaft 26 equipped with permanent magnets is disposedradially inside the stator 20. This outer rotor shaft is also referredto as a “rotor package”. The outer rotor shaft 26 comprises on its innerside an elastic region which is plugged onto an inner rotor shaft 27. Onthe inner rotor shaft 27 an eccentrically arranged oval cam disk 28 isformed on the output side, which is shown in more detail in FIG. 3.

The inner rotor shaft 27 is supported in the motor housing 22 on thedrive side to the outside by a drive-side rotor ball bearing 29. Namely,an outer ring of the drive-side rotor ball bearing 29 is disposed in acylindrical recess of the motor housing 22.

Furthermore, the inner rotor shaft 27 is mounted on the output side inan output-side rotor ball bearing 30 radially outwardly in an inner gear7. A hollow shaft of the inner gear 7 is connected integrally with anannular portion of the inner gear 7 on the output side, which comprisesan external toothing 5.

The hollow shaft of the inner gear 7 is in turn mounted radiallyoutwardly via an inner gear ball bearing 31 on a housing cover 32 whichis screwed to the motor housing 22 by screws 33. The inner gear ballbearing 31 is offset from the output side rotor ball bearing 30 in theaxial direction to the output side and is offset in the radial directionto the outside. In addition, the inner gear ball bearing 31 overlapswith the output side rotor ball bearing 30 in the axial direction.

On the cam disk 28 of the inner rotor shaft 27, a flexible ball bearingor a thin section ball bearing 33 is clamped, in which an inner ring andan outer ring are deformable. A pin retaining ring 103 with pins 101bears on the flexible ball bearing 33, wherein the pins 101 are held incylindrical recesses on the inside of a pin retaining ring. According tothe embodiment of FIG. 1, the pins 101 are connected to each other. Thepins 101 of the pin ring 103 protrude the flexible ball bearing 33 andthe pin retaining ring 103 in axial direction on both sides. For thesake of simplicity, the pin ring assembly comprising the pin retainingring 103 and the pins 101 will also be referred to below as pin ring102.

The cam disk 28 and the flexible ball bearing 33 together form atransmitter arrangement, which converts a torque into a radial force.Instead of a flexible ball bearing with flexible inner and outer ring, awire-race bearing or a flexible ball bearing without outer ring, or adifferent kind of flexible rolling bearing can be used.

The housing cover 32 is screwed to the motor housing with fasteningscrews 34 on the output side of the motor housing. Furthermore, a driveside outer gear 8′ and an output side outer gear 8 are inserted into asupport ring 36 and screwed to the support ring 36 by the screws 34. Thesupport ring 36 is divided in the axial direction into two mutuallymirror-symmetrical halves, which together form a raceway 67 for the pinretaining ring 103.

The drive side outer gear 8′ and the output side outer gear 8 arearranged in the axial direction outside the cam disk 28 and the flexibleball bearing 33. In the radial direction, the drive side outer gear 8′is opposite to the areas of the pins 101 which protrude the pinretaining ring 103 on the drive side in the axial direction. In theradial direction, the output side outer gear 8 is opposite to the areasof the pins 101 which protrude the pin retaining ring 103 on the outputside in the axial direction.

On the drive side, a drive side spacer disk 37 is arranged in the motorhousing 22 such that it faces the drive side end faces of the pins 101in the axial direction. Similarly, on the output side, an output sidespacer disk 38 is arranged in the motor housing 22 such that it facesthe output side end surfaces of the pins 101 in the axial direction.

An output shaft 39 is disposed radially inside the hollow shaft of theinner gear 7, wherein a transmission freewheel 40 is disposed betweenthe hollow shaft of the inner gear 7 and the output shaft 39. The outputshaft 39 is mounted radially outwardly in an output side output ballbearing 41, which is inserted into a cylindrical recess or shoulder ofthe housing cover 32. The output side region of the output shaft 39protrudes the housing cover 32 in the axial direction. A chainringadapter 43 is mounted on the output shaft 39 and is held via a roundoutput nut 44 which is screwed into an internal thread of the outputshaft 39.

The pedal shaft 35 is disposed partially in the interior of the rotorshaft 27 and partially in the interior of the output shaft 39 andconcentric with the rotor shaft 27 and the output shaft 39. The pedalshaft 35 is mounted via a drive side pedal shaft ball bearing 45radially outward in a load cell 47, which in turn is inserted into themotor housing 22. A printed circuit board 48, also referred to as a “PCBforce sensor”, with evaluation electronics is fastened to the load cell47 and a connector of the printed circuit board 48 is guided radiallyoutwards via a ribbon cable 63 and connected to the electronics on theprinted circuit board 23.

The load cell 47 comprises strain gauges that generate electricalsignals that correspond to a deformation of the load cell 47 serving asa suspension of the drive side pedal shaft ball bearing. The load cell47 comprises four webs, which are arranged at 45 degrees spacing in thecircumferential direction and which are connected to a ring in which theouter ring of the output side bottom bracket ball bearing is inserted.On each of these webs in each case a strain gauge is applied, which iselectrically connected to the annular printed circuit board 22.

In the drive side direction of the output side pedal shaft ball bearing46, a pedal shaft freewheel 49 is arranged between the pedal shaft 35and the output shaft 39. The transmission freewheel 40, the hollow shaftof the inner gear 7 and the inner gear ball bearing 31 follow radiallyoutwardly. Instead of a single pedal shaft freewheel 49, two adjacentfreewheels or a single freewheel and an adjacent rolling bearing, suchas a needle bearing can be installed.

On the output side of the output side pedal shaft ball bearing 46, aninner shaft seal ring 50 is inserted between the pedal shaft 35 and theoutput shaft 39 opposite the output side pedal shaft ball bearing 46.Furthermore, an outer shaft seal ring 51 is disposed opposite the outputside output ball bearing 46 between the output shaft 39 and the housingcover 32. Another shaft seal ring 52 is arranged on the drive sidebetween the cooling cover 24 and the pedal shaft 35. An O-ring 42 isinserted radially outward between the output side outer gear 8 and thehousing cover 32.

In operation, an input torque is transmitted via the stator 20 byelectromagnetic force action to the outer rotor shaft 26 and from thereto the inner rotor shaft 27, which is converted by the cam disk 28 andthe flexible ball bearing 33 into a radial force. This radial force isconverted at the tooth flanks of the internal toothings 6, 6′ of theouter gears 8, 8′ and the external toothing 5 of the inner gear 7 intoan output torque, wherein the inner gear 7 is driven, and the outergears 8, 8′ are fixed to the housing. The output torque is larger thanthe input torque by the reduction gear ratio.

The inner toothing, which is formed by the external toothing 5 of theinner gear 7, lies opposite the internal toothing 6 of the output sideouter gear 8, and thereby provides the output torque, in particular bythose pins 101 which abut on both the external toothing 5 and theinternal toothings 6, 6′.

FIG. 2 shows an exploded view of the transmission of FIG. 1, in which,viewed from the drive side to the output side, the drive side rotor ballbearing 29, the inner rotor shaft 27, the cam disk 28, the output siderotor ball bearing 30, the second outer gear 8′, the flexible ballbearing 33, the pin retaining ring 103 with the pins 101, the supportring 36, the first outer gear 8, the inner gear 7 with the inner gearhollow shaft and the inner gear ball bearing 31 are shown.

The outer gears 8, 8′ each comprise journals 53 which project radiallyoutwards from the respective outer gear 8, 8′ and are distributed atregular intervals over the circumference of the outer gears 8, 8′. Thesupport ring 36 comprises radial slots 54 on radially opposite sidescorresponding to the journals 53 and distributed on the circumference.In addition, screw holes 55 are provided for fastening the outer gears,which, in the embodiment of FIG. 2, are located partially in the outergears 8, 8′ and partially in the support ring 36.

Between the drive side rotor ball bearing 29 and the motor housing 22, awave spring 61 is arranged on the drive side of the drive side rotorball bearing 29 and a spacer ring 62 is arranged between the drive siderotor ball bearing 29 and the outer rotor shaft 26 on the output side ofthe drive side rotor ball bearing 29.

In particular, the support ring 36 may be made of aluminum and the outergears 8, 8′ of plastic, such as polyamide 66 (PA66), wherein the supportring 36 may be made in particular by aluminum die-cast, and the outergears may be made in particular by a plastic injection molding.Furthermore, the inner rotor shaft 27 may be made of aluminum.

As shown in FIG. 1, the screws 34 extend through the transmission cover32, the first outer gear 8, the second outer gear 8′, and the supportring 36 into the motor housing 22 in the assembled state.

FIG. 3 shows a drive-side part of an exploded view of the transmissionof FIG. 1, in which, viewed from the drive side, the cooling cover 24,the printed circuit board 23, the motor housing 22 with the stator 20and the coil 21 or with the coils 21, the printed circuit board 48, theload cell 47, the wave spring 61, the drive side spacer disk 37, thesecond outer gear 8′ with the internal toothing 6′, the drive side rotorball bearings 29, the spacer ring 62, the outer rotor shaft 26, and apart of the inner rotor shaft 27 are shown.

FIG. 3 shows an output-side part of an exploded view of the transmissionof FIG. 1, in which, viewed from the drive side, the inner rotor shaft27 with the cam disk 28, the output side rotor ball bearing 30, the pinring 102 with the pin retaining ring 103 and the pins 101, the supportring 36, the output side outer gear 8, the inner gear 7 with theexternal toothing 5, the output shaft 37, the transmission freewheel 40,the O-ring, the output side spacer disk 38, the inner gear ball bearing31, the housing cover 32 with the fastening screws 34 and the shaftsealing ring 50 are shown.

As shown in FIG. 3, the transmission freewheel 40 comprises coil springscylindrical rollers 64 which are arranged in a clamping body cage 65.For pressing the cylindrical rollers 64, coil springs 66 are provided inthe clamping body cage 65.

FIG. 4 shows an exploded view of an inner assembly of the transmissionof FIG. 1, and in particular the pedal shaft freewheel 49. In detail,FIG. 4, seen from the drive side, shows a sensor ring 68, the drive sidepedal shaft ball bearing 45, the pedal shaft 35, the pedal shaftfreewheel 49, the output side pedal shaft ball bearing 46, a spacer 69,a chainring adapter 43, a wave spring 70, an inner shaft seal 50, anO-ring 76 and the output nut 44.

As shown in FIG. 4, the pedal shaft freewheel 49 comprises blade-shapedsteps 71 which are formed on the pedal shaft 35. Pawls 73 and coilsprings 72 are arranged between the steps 71. At both ends of the pedalshaft, fastening regions 74, 75 for pedal cranks are arranged, which areshown in FIG. 1 in cross section.

FIG. 5 shows a perspective view of an inner assembly of the transmissionof FIG. 1 in the assembled state. For the sake of clarity, the innergear 7 and the flexible ball bearing 33 have been omitted in FIG. 5.

FIG. 6 shows a pedal shaft assembly 80 or a pedal shaft unit 80 of thetransmission 10 of FIG. 1, wherein the drive side is right and theoutput side left. The pedal shaft assembly 80 comprises a freewheelassembly 81 and a sensor assembly 82. The viewing direction of the viewof FIG. 6 is against the direction of driving in the installed state.

FIG. 7 shows a cross section through the freewheel assembly 81 of FIG. 6along the intersection line D-D, which is viewed from the output side,wherein in the background the flat cable 63 belonging to the sensorassembly 82 is visible. Furthermore, on the right side of the load cell47 screw heads of the screws can be seen, with which the annular printedcircuit board 23 is screwed to the load cell 47.

The freewheel assembly 81 comprises the pedal shaft freewheel 49 and thetransmission freewheel 40, wherein the output shaft 39 is additionallyconfigured as outer ring of the pedal shaft freewheel 49 and as innerring of the transmission freewheel 40. As shown in FIG. 7, the pedalshaft freewheel 49 is configured as pawl freewheel while thetransmission freewheel 40 is configured as a clamp roller freewheel withcylindrical clamp rollers 64. The inner circumference of the outputshaft 39 forms a serrated staircase whose steps form stops for the pawls73, which are pressed by the coil springs 72 to the inner circumferenceof the output shaft 39.

Similarly, the outer circumference of the output shaft 39 forms aserrated staircase, also referred to as a “star”, whose steps form stopsfor the cylindrical rollers 64 of the transmission freewheel 40. Thefreewheel cage 65 and thus the cylindrical rollers 64 are pressed by thecoil springs 66, which are arranged between the outer circumference ofthe output shaft 39 and the freewheel cage 65, to the outercircumference of the drive shaft 39.

The steps of the drive shaft 39 are arranged such that a freewheelingdirection of the pedal shaft freewheel 49 and a freewheeling directionof the transmission freewheel 40 run in the counterclockwise directionin the view of FIG. 7. This freewheeling direction is in each casereversed to the respective drive direction or locking direction.

Thus, the pedal shaft 35 can drive the output shaft 39 in a clockwisedirection, as long as the output shaft 39 does not move faster than thepedal shaft 35 and the outer ring of the transmission freewheel 40,which is formed by an inner portion of the inner gear 5, can drive theoutput shaft 39 in the clockwise direction as long as the output shaft39 does not move faster than the outer ring. The drive and lockingdirections, respectively, of the freewheels 40, 49 are indicated in FIG.7 in each case by arrows.

FIG. 10 is a simplified cross-sectional view illustrating the load cell47 used in the transmission of FIG. 1 and the use of an obliquelymounted bearing. For simplicity, the electric motor and the drive trainof the electric motor are omitted in this illustration, so that thepedal shaft 35 is mounted directly on a housing cover 32′.

According to FIG. 10, the drive-side pedal shaft ball bearing 45 and theoutput side pedal shaft ball bearing 46 are each configured as anangular contact ball bearing, which can receive axial forces to acertain extent. Instead of ball bearings, other types of bearings can beprovided, which can also receive axial forces in addition to radialforces, such as taper roller bearings. However, this is usually moreexpensive than the use of ball bearings.

The load cell 47, which is mounted on the pedal shaft 35 via the bearing45, comprises support lugs 91 and measuring lugs 90. The support lugs 91are supported in the axial direction on an outer ring of the ballbearing and the measuring lugs 90 are supported in the radial directionon the outer ring. Strain gauges are attached to at least two of the 90measuring lugs. The inner ring of the ball bearing 45 abuts towards thecenter of the pedal shaft 35 on a shoulder or on a step of the pedalshaft 35.

The inner ring of the ball bearing 46 abuts towards the center of thepedal shaft 35 via a wave spring 70 on a shoulder of the pedal shaft 35.The outer ring of the ball bearing 46 abuts on the transmission cover44′ via a spacer disk.

In operation, a rider transmits radial forces via a pedal crank to thepedal shaft 35. These radial forces are received by the measuring lugs90 and lead to a deformation of the measuring lugs 90, which isdetermined by the strain gauges. In contrast, axial forces on the outerring of the ball bearing 45 do not lead to a deformation of the secondlugs. Instead, the axial forces are absorbed by the support lugs 91 ofthe load cell 47, whereby the ball bearing 45 is held laterally.

Compared to the force of the rider on the load cell, any existing torquethrough an auxiliary drive does not lead or only slightly leads to adeformation of the second lugs. Thus, the contribution of the rider canbe determined separately. In addition, an angular position sensor can beprovided, with the help of which the position of the pedal cranks can bedetermined and thereby the lever arm of the pedals. By a suitablecalculation model, which is implemented by a stored program and/or acircuit, the torque provided by the rider can be reconstructed from themeasured radial forces.

FIG. 10 illustrates a transmission of axial forces in the housingaccording to the mounting concept of the transmission of FIGS. 1 to 9.

As shown at the bottom right in FIG. 10, an output side axial force ofthe pedal shaft 35 is transmitted via a shoulder of the pedal shaft 35,the output side pedal shaft ball bearing 46, the wave spring 70, theoutput nut 44, the internal thread of the output hollow shaft 39, theoutput hollow shaft 39, a step of the output hollow shaft 39, the outputball bearing 41, the housing cover 32, the screws 34 and the screwthread 60 into the housing 22.

A drive side axial force of the pedal shaft 35 is transmitted via adrive side step 123 of the pedal shaft 35, the drive side pedal shaftball bearing 45, axial support lugs of the load cell 47 and a mountingring of the load cell 47 into the housing 22.

As shown at the top right of FIG. 10, an output side axial force of therotor shaft 28 is transmitted via the output side rotor ball bearing 30,the inner gear 7 and the inner gear ball bearing 31 into the housingcover. From there, the axial force is transmitted via the screws 34 intothe housing 22, as shown at the bottom right in FIG. 10.

A drive side axial force of the rotor shaft 27 is transmitted via ashoulder 9 of the rotor shaft 27, the outer rotor shaft 26, the spacerring 62, the drive-side rotor ball bearing 29 and the wave spring 61into the housing 22.

Furthermore, an axial force on the ball bearing 33 arranged on the camdisk is transmitted via the inner gear 7 and the ball bearing 31 intothe transmission housing 22. The counterforce thereto is transmitted viaa step 9 of the cam disk 28 into the rotor shaft 27 and from there viathe above-described path into the housing 22.

The inner gear 7 tapers radially inwardly of an outer circumference, sothat only the outer ring of the ball bearing 33 abuts against the innergear 7, which moves essentially synchronously with the pin ring 101 andwith the inner gear 7, and not the inner ring of the ball bearing 33,which rotates much faster than the inner gear 7. The width of the wavespring 70 is adjusted such that the chainring adapter 43 does not abutagainst the housing cover 32.

FIGS. 11 to 20 show a harmonic pin ring transmission with an eccentriccam disk, wherein the eccentric cam disk is an eccentrically arrangedcircular disk.

FIG. 11 shows a cross-sectional view of a harmonic pin ring transmission10. The cross-sectional plane A-A of FIG. 11 is indicated in FIG. 19. InFIG. 11, the left side corresponds to a drive side and the right sidecorresponds to an output side of the harmonic pin ring gear 10.

A stator 20 of a stator assembly of the harmonic pin ring transmission10 is disposed in a motor housing 22. The stator 20 comprises threeseparate coils 21 for connecting to three phases of a three-phaseinverter. The three-phase inverter is configured as a power electronics,which is arranged on a printed circuit board 23, wherein the printedcircuit board 23 is arranged in a cooling cover 24, which is mounted onthe motor housing 22 on the drive side. The three coils of the stator 20are connected to the three-phase inverter via three terminals 25, one ofwhich is shown in FIG. 11.

The motor housing 22 and the cooling cover 24 each have cylindricalprotruding portions which surround a pedal shaft 35 extending centrallythrough the motor housing 22. The printed circuit board 23 is formed asa perforated disk, which is located outside of the protruding portions,so that the electronics located thereon is sealed against oil and greaseof the transmission.

An outer rotor shaft 26 equipped with permanent magnets is disposedradially inside the stator 20. This outer rotor shaft 26 is alsoreferred to as a “rotor package”. The outer rotor shaft 26 comprises onits inner side an elastic region which is plugged onto an inner rotorshaft 27. On the inner rotor shaft 27, an eccentrically arrangedcircular eccentric disk 28′ is formed on the output side, which is shownin more detail in FIG. 12.

On the drive side, the inner rotor shaft 17 is supported in the motorhousing 22 to the outside by a drive-side rotor ball bearing 29. Namely,an outer ring of the drive side rotor ball bearing 29 is disposed in acylindrical recess of the motor housing 22.

Furthermore, the inner rotor shaft 27 is mounted on the output side inan output side rotor ball bearing 30 radially outwardly in an inner gear7. A hollow shaft of the inner gear 7 is connected integrally with anannular portion of the inner gear 7 on the output side, which comprisesan external toothing 5.

The hollow shaft of the inner gear 7 is in turn mounted radiallyoutwardly via an inner gear ball bearing 31 on a housing cover 32 whichis screwed to the motor housing 22 by screws 34. The inner gear ballbearing 31 is offset from the output side rotor ball bearing 30 in theaxial direction to the output side and is offset in the radial directionto the outside. In addition, the inner gear ball bearing 31 overlapswith the output side rotor ball bearing 30 in the axial direction.

On the eccentric circular disk 28 of the inner rotor shaft 27, a ballbearing 33 is clamped. A pin retaining ring 103 with pins 101 bears onthe ball bearing 33, wherein the pins 101 are held in cylindricalrecesses on the inside of a pin retaining ring.

The pin ring assembly of pins 101 and pin retaining ring 103 is shown inmore detail in FIGS. 15 to 17. According to the embodiment of FIGS. 15to 17, the pin ring 103, unlike in the embodiment of FIGS. 1 to 10, isformed from one piece and comprises areas corresponding to the pins 101and the pin retaining ring 103. The one-piece embodiment of the pin ring103 is particularly well suited for a transmission in which the pin ringis not or only slightly deformed. This is the case, for example, in atransmission with an eccentrically arranged circular disk, which isshown in FIGS. 11 to 20, 61 to 66, 67 and 72. The pins 101 of the pinring 103 protrude in the axial direction on both sides the flexible ballbearing 33 and the pin retaining ring 103.

The eccentric disc 28′ and the ball bearing 33 together form atransmitter assembly which converts a rotary motion into a radial motionwhich is transmitted by the ball bearing 33 and the pin ring 103 to thepins 101, and which is then converted back again into a rotary motion bythe engagement of the pins 101 in the outer gears 8, 8′.

The housing cover 32 is screwed to the motor housing with fasteningscrews 34 on the output side of the motor housing. Furthermore, a driveside outer gear 8′ and an output side outer gear 8 are inserted into asupport ring 36 and screwed to the support ring 36 by the screws 34. AnO-ring 42 is inserted radially outward between the output side outergear 8 and the housing cover.

The drive side outer gear 8′ and the output side outer gear 8 arearranged in the axial direction outside the eccentric disk 28′ and theball bearing 33. In the radial direction, the drive side outer gear 8′is opposite to the areas of the pins 101 which protrude the pinretaining ring 103 on the drive side in the axial direction. In theradial direction, the output side outer gear 8 is opposite to the areasof the pins 101 which protrude the pin retaining ring 103 on the outputside in the axial direction.

On the drive side, a drive side spacer disk 37 is arranged in the motorhousing 22 such that it faces the drive side end faces of the pins 101in the axial direction. Similarly, on the output side, an output sidespacer disk 38 is arranged in the motor housing 22 such that it facesthe output side end surfaces of the pins 101 in the axial direction.

An output shaft 39 is disposed radially inside the hollow shaft of theinner gear 7, wherein a transmission freewheel 40 is disposed betweenthe hollow shaft of the inner gear 7 and the output shaft 39. The outputshaft 39 is mounted radially outwardly in an output side output ballbearing 41, which is inserted into a cylindrical recess or shoulder ofthe housing cover 32. The output side region of the output shaft 39protrudes the housing cover 32 in the axial direction. A chainringadapter 43 is mounted on the output shaft 39 and is held via atransmission cover 44 which is screwed into an internal thread of theoutput shaft 39.

The above-mentioned pedal shaft 35 is disposed partially in the interiorof the rotor shaft 27 and partially in the interior of the output shaft39 and concentric with the rotor shaft 27 and the output shaft 39. Thepedal shaft 35 is mounted via a drive side pedal shaft ball bearing 45radially outward in a load cell 47, which in turn is inserted into themotor housing 22. A printed circuit board 48 with a force sensor isfastened to the load cell 47 and a connector of the printed circuitboard 48 is guided radially outwards via a ribbon cable 63 and connectedto the electronics on the printed circuit board 23.

In the drive side direction of the output side pedal shaft ball bearing46, a pedal shaft freewheel 49 is arranged between the pedal shaft 35and the output shaft 39. The transmission freewheel 40, the hollow shaftof the inner gear 7 and the inner gear ball bearing 31 follow radiallyoutwardly. Instead of a single pedal shaft freewheel 49, two adjacentfreewheels or a single freewheel and an adjacent rolling bearing, suchas a needle bearing can be installed.

On the output side of the output side pedal shaft ball bearing 46, aninner shaft seal ring 50 is inserted between the pedal shaft 35 and theoutput shaft 39 opposite the output side pedal shaft ball bearing 46.Furthermore, an outer shaft seal ring 51 is disposed opposite the outputside output ball bearing 46 between the output shaft 39 and the housingcover 32. Another shaft seal ring 52 is arranged on the drive sidebetween the cooling cover 24 and the pedal shaft 35.

In operation, an input torque is transmitted via the stator 20 byelectromagnetic force action to the outer rotor shaft 26 and from thereto the inner rotor shaft 27, which is converted by the eccentric disk28′ and the ball bearing 33 into a radial force. This radial force isconverted at the tooth flanks internal toothings 6, 6′ of the outergears 8, 8′ and the external toothing 5 of the inner gear 7 into anoutput torque, wherein the inner gear 7 is driven, and the outer gears8, 8′ are fixed to the housing. The output torque is larger than theinput torque by the reduction gear ratio.

The inner toothing, which is formed by the external toothing 5 of theinner gear 7, lies opposite the internal toothing 6 of the output sideouter gear 8, and thereby provides the output torque, in particular bythose pins 101 which abut on both to the external toothing 5 and to theinternal toothings 6, 6′.

FIG. 12 shows an exploded view of the transmission of FIG. 11, in which,viewed from the drive side to the output side, the drive side rotor ballbearing 29, the inner rotor shaft 27, the eccentric disk 28 ‘, theoutput side rotor ball bearing 30, the second outer gear 8’, the ballbearing 33, the pin retaining ring 103 with the pins 101, the supportring 36, the first outer gear 8, the inner gear 7 with the inner gearhollow shaft and the inner gear ball bearing 31 are shown.

The outer gears 8, 8′ each comprise journals 53 which project radiallyoutwards from the respective outer gear 8, 8′ and are distributed atregular intervals over the circumference of the outer gears 8, 8′. Thesupport ring 36 comprises radial slots 54 on radially opposite sidescorresponding to the journals 53 and distributed on the circumference.In addition, screw holes 55 are provided for fastening the outer gears,which, in the embodiment of FIG. 12, are located partially in the outergears 8, 8′ and partially in the support ring 36.

Between the drive side rotor ball bearing 29 and the motor housing 22, awave spring 61 is arranged on the drive side of the drive side rotorball bearing 29 and a spacer ring 62 is arranged between the drive siderotor ball bearing 29 and the outer rotor shaft 26 on the output side ofthe drive side rotor ball bearing 29.

In particular, the support ring 36 may be made of aluminum and the outergears 8, 8′ of plastic, such as polyamide 66 (PA66), wherein the supportring 36 may be made in particular by aluminum die-cast, and the outergears may be made in particular by a plastic injection molding.Furthermore, the inner rotor shaft 27 may be made of aluminum.

As shown in FIG. 11, the screws 34 extend through the transmission cover32, the first outer gear 8, the second outer gear 8′, and the supportring 36 into the motor housing 22 in the assembled state.

FIG. 13 shows a side view, viewed from the output side, of an assembledarrangement of inner rotor shaft 27, inner gear 7, pins 101 and outergear 8. As can be seen in FIG. 13, at a predetermined position of theeccentric disc 28′ at a first position, the pins 101 completely abutagainst the outer internal toothing 6 and at a second position which isopposite the first position, the pins 101 completely abut against theinner external toothing 5.

FIG. 14 shows a cross-sectional view through the arrangement of FIG. 13along the intersection line E-E. As shown in the cross-sectional view ofFIG. 14, the eccentric disc 28′ comprises a step 11 on the drive side,through which an inner ring of the ball bearing 33 is held from thedrive side.

FIG. 15 shows an output side view of a pin ring 102 for use in thetransmission of FIG. 11. FIGS. 16 and 17 show detailed views of the pinring 102 of FIG. 15, and FIG. 18 shows a drive side view of the pin ring102 of FIG. 15.

The pin ring 102 comprises a pin retaining ring 103 and pins 101. Thepin ring 102 is integrally formed, wherein the pin retaining ring 103and the pins 101 are each formed by portions of the pin ring 102, whichis formed in one piece. The pins 101 are respectively formed differentlyon an output side and on an opposite drive side.

Namely, the pins on the output side of the pin ring 102, at which thepins 101 engage both the inner gear 7 and the outer gear 8, are lensshaped, wherein two radially opposite regions each comprise a circularsegment shaped boundary and two peripherally opposite regions taper.This is best seen in the side view of FIG. 17, which is marked in FIG.15 with “G”.

On the other hand, the pins 101 on the drive side where the pins 101engage only the second outer gear 8′ are formed as “half pins” whichrespectively form teeth of a comb extending along the circumference ofthe pin ring 102 as shown in the cross sectional view of FIG. 16 and inthe view of FIG. 18.

In the side view of FIG. 17, a first region 255 of a pin 101 is shown,which transmits torque to the inner gear 7, and a second, radiallyopposite region 56, which transmits torque to the outer gear 8.Tangential transitions between the pins 101 occur due to the regions ofthe pins 101 tapering along the circumferential direction. As a result,on the one hand, an economical CNC production of the pins 101 can beprovided and, on the other hand, a higher tangential rigidity of thepins 101 can be achieved.

According to another manufacturing method, the pin ring is made bytransfer molding. According to this method, first a hollow ring isproduced, from which the pins 101 are then formed by parting off ormilling off. In a shaping by milling, a rotating milling machine can beused, which moves on a circular line, wherein the pins are formed in theradial direction first from the inside, then from the outside.

The transfer molding is further defined, for example, in the standardsDIN 8582 and DIN 8583. According to a special method, a workpiece to beformed is first applied to a mandrel. The workpiece to be formed is thendeformed from the outside by transfer molding wheels, which are drivenby a traverse drive system in a controlled manner, press against theworkpiece and thin the material in a controlled manner.

Furthermore, on the output side of the pins 101 a step 57 is formedradially inwardly. As can be seen in FIG. 11, this step 57 bears on theoutput side on the outer ring of the ball bearing 33 in the installedstate of the pin ring 102.

As shown in the drive-side view of the pin ring 102 of FIG. 8, a combwith round recesses 58 and intermediate plateaus 59 is formed on thedrive side of the pin ring 102. The intermediate plateaus 59 form anextension of the pins 101 shown in detail in FIG. 17 and can thus beregarded as the drive side region of the pins 101.

FIG. 19 shows an output side view of the assembled transmission of FIG.11, in which the cross-sectional plane of FIG. 11 is labeled “A”.

FIG. 20 illustrates a transmission of axial forces into the housingaccording to the mounting concept of the transmission of FIGS. 11 to 19.

As shown at the bottom right in FIG. 20, an output side axial force ofthe pedal shaft 35 is transmitted via a shoulder of the pedal shaft 35,the output side pedal shaft ball bearing 46, the wave spring 70, theoutput nut 44, the internal thread of the output hollow shaft 39, theoutput hollow shaft 39, a step of the output hollow shaft 39, the outputball bearing 41, the housing cover 32, the screws 34 and the screwthread 69 into the housing 22.

A drive side axial force of the pedal shaft 35 is transmitted via adrive side step 123 of the pedal shaft 35, the drive side pedal shaftball bearing 45, axial support lugs of the load cell 47 and a mountingring of the load cell 47 into the housing 22.

As shown at the top right of FIG. 10, an output side axial force of therotor shaft 27 is transmitted via the output side rotor ball bearing 30,the inner gear 7 and the inner gear ball bearing 31 into the housingcover. From there, the axial force is transmitted via the screws 34 intothe housing 22, as shown at the bottom right in FIG. 20.

A drive side axial force of the rotor shaft 27 is transmitted via adrive side step of the rotor shaft 27, the outer rotor shaft 26, thespacer ring 62, the drive-side rotor ball bearing 29 and the wave spring61 into the housing 22.

Furthermore, an axial force is transmitted via a shoulder 9 of the camdisk 28 to the ball bearing 33 via a shoulder 10 of the pin ring 100,the inner gear 7 and the ball bearing 31 into the transmission housing22. The counterforce thereto is transmitted via the step 9 of the camdisk 28 into the rotor shaft 28 and from there via the above-describedpath into the housing 22.

The inner gear 7 tapers radially inwardly of an outer circumference, sothat only the outer ring of the ball bearing 33 abuts against the innergear 7, which moves essentially synchronously with the pin ring 101 andwith the inner gear 7, and not the inner ring of the ball bearing 33,which rotates much faster than the inner gear 7. The width of the wavespring 70 is adjusted such that the chainring adapter 43 does not abutagainst the housing cover 32.

The embodiments of FIGS. 21 to 44 disclose devices and methods formeasuring a torque applied to a pedal shaft of a bicycle assisted by anelectric drive.

By measuring the torque on a pedal shaft, the support of an electricmotor for an electric bicycle can be adjusted. For this purpose, variousmethods are known. For example, the torque can be determined withoutcontact by means of a magnetic measurement of the twisting of the pedalshaft. Another method determines a mechanical deformation of the bottombracket ball bearing suspension. A corresponding device is disclosed forexample in the patent application DE 102013220871 A1. According to thisdevice, a single electromagnetic or mechanical sensor is used todetermine a horizontal deflection or deformation of the bottom bracketball bearing suspension.

The present description discloses a load cell for determining a radialforce acting on a crankshaft with a receiving sleeve for receiving aring of a bearing and a fastening ring for attaching the load cell in atransmission housing. Axial support areas are provided on the fasteningring for axially supporting the outer ring of the first bearing.Measuring regions that connect the receiving sleeve with the fasteningring are provided for receiving radial forces of the receiving sleeve.

Strain sensors are attached to at least two of the measuring regions.The strain sensors may be configured in particular as strain gauges,wherein the strain gauges may be glued to a surface of the measuringregions.

The axial support areas and the measuring regions may in particular beconfigured as lugs or fingers. Furthermore, the measuring regions may beconfigured as angular lugs. The radial force acting on the crankshaft isindirectly determined by a radial force acting on the receiving sleeve,which in turn is transmitted via a bearing from the crankshaft to thereceiving sleeve. The crankshaft may in particular be a pedal shaft.

The load cell can be used in particular in electric bicycles forregulating a motor support, wherein a pedal shaft is supported via theload cell in the axial direction. In a broader sense, the load cell canbe used anywhere where a radial force is converted into a rotationalmovement of the shaft on a shaft, in particular in a bicycle pedal shaftor a piston which is driven by an internal combustion engine or anotherdrive. In this case, the measurement of the radial forces via the loadcell may allow a calculation of the torque applied to the shaft.However, the radial forces can also be used to generate control signalswithout the intermediate step of a torque calculation.

If there is no conversion of radial forces, such as in a geartransmission, a load cell as described herein can also be used todetermine an imbalance on the shaft. The load cell is advantageous inparticular in the case of a mechanical drive, such as in a pedal shaft,since in this case the applied torque can not readily be determined viaa motor power, which is easily possible, for example, with an electricmotor.

In a load cell according to the present description, radial and axialforces are decoupled by the fact that axial forces are received by theaxial support areas or lugs and radial forces are received by themeasuring regions or lugs. Thereby, the load cell can be used in anobliquely mounted bearing, such as in an O-arrangement or anX-arrangement. Additionally, decoupling the radial from the axial forcesmay allow a more accurate measurement of the radial force component, andthus a more accurate determination of the torque applied to the shaft.

A load cell according to the present description is simple and requiresonly little space in a transmission, in particular in the axialdirection. The total width of the load cell is essentially determined bythe width of the bearing and by the width of the axially adjoiningfastening region. Therefore, the load cell can be made relatively narrowin the axial direction, for example, it can only have about twice thewidth of a rolling bearing or less.

The absorption of the axial forces by the load cell is advantageouslycarried out on the stationary outer ring of the ball bearing and not onthe movable inner ring or on the shaft, where the relative movementwould lead to friction losses. Thus, the axial support areasadvantageously only touch the outer ring of the bearing. The supportfunction of the axial support areas can be achieved in particular inthat the axial support areas or lugs protrude radially inwardly over aninner surface of the receiving sleeve, it being sufficient if the axialsupport areas only protrude inwardly as it corresponds to the width of abearing ring. The axial support areas may also have a profile foralignment with the bearing ring.

The measuring lugs are expediently configured as angle brackets, whereinan axial region of the angle bracket connected to the receiving sleeveextends in the axial direction and a radial region of the angle bracketconnected to the fastening region extends in the radial direction. Assuch, the angle bracket forms a lever arm which causes bending of theradial portion of the measuring lugs relative to the radial orientation.Due to the bending, a surface of the measuring lugs is stretched orcompressed and an axially opposite surface is compressed or stretched.This compression or stretching is detected by the strain gauges appliedto one of the two surfaces of the radial region.

The two regions of the angle bracket formed by the measuring lug canalso be slightly inclined with respect to the axial or the radialorientation, for example, to allow greater deformation. Advantageously,the measuring lugs are sufficiently thin in the axial direction to allowa good deformation. In this case, the second region may be made thinnerthan the first region. Furthermore, the second region near the top ofthe angle bracket may be reinforced to avoid deformation near the top ofthe angle bracket.

The axial region of the measuring lug is advantageously configured suchthat it does not protrude beyond the surface of the receiving sleeve,which rests on the rolling bearing. This avoids that a ring of a rollingbearing on which the receiving sleeve rests, abuts the measuring lug andaxial forces are transmitted via the measuring lug. In particular, theaxial region may be flush with a surface of the receiving sleeve, thesurface being a cylindrical inner surface of the receiving sleeve.

According to an exemplary embodiment, the measuring regions comprisemeasuring lugs formed as angle brackets or are configured asangle-shaped measuring lugs. According to a further exemplaryembodiment, the measuring lugs comprise a radial region and an axialregion adjoining the radial region. In this case, the radial region isconnected to the fastening ring and the axial region is connected to thereceiving sleeve, wherein the radial region is arranged to the axialregion at an angle of approximately 90 degrees.

According to an advantageous embodiment, four measuring regions ormeasuring lugs are provided so that the shaft can be supported in fourdirections. In particular, the measuring lugs can be arranged at adistance of 90 degrees, so that the directions are perpendicular to oneanother. The measuring lugs are at the same time supporting lugs forradial support. The term “measuring lug” means that the deformation ofthe measuring lug can be used to measure forces that act on thereceiving sleeve in a radial direction.

According to a further embodiment, an angular extent of the measuringlug in the circumferential direction can be suitably selected in orderto detect a force in a defined direction and to enable a gooddeformation. For example, the angular extent of the measuring lug alongthe circumferential direction may be limited to 30 degrees or less, or25 degrees or less, wherein 90 degrees correspond to a right angle.

According to an embodiment, the load cell comprises four measuringregions, which are arranged at a distance of 90 degrees. Thus, radialforces can be measured in four mutually perpendicular directions whenall four measuring regions or measuring lugs are equipped with strainsensors.

In this case, at least one, several or all of the strain sensors may beformed as strain gauges, which are inexpensive and robust, in particularcompared with an optical strain measurement.

In particular, a strain sensor may be attached to each of the measuringranges or measuring lugs to measure radial forces in as many directions.For ease of attaching the strain gauges, the surfaces of the measuringranges to which the strain gauges are attached may have recessed areasor pockets for attachment of the strain sensors or strain gauges.

The strain gauges may be arranged close to the fastening region, forexample to keep the electrical connection to a printed circuit boardattached to the transmission housing as short as possible or becausethere occurs a greater deformation.

According to a further embodiment, the axial support areas or lugs areseparated from the measuring regions by radial slots. Furthermore, theaxial support areas or lugs are separated from the receiving sleeve by acircumferential slot. Among other things, the axial support areas andthe measuring regions thereby can easily be formed from a workpiece. Theload cell may however also be cast in this form. Conveniently, theradial or circumferential slots are straight slots in radial orcircumferential direction, respectively. However, the shape may alsodiffer from a straight line.

The direction of the radial slots may deviate from a radial direction,for example, to cause the measuring lug to taper more inwardly. Forexample, the orientation of the radial slots can be inclined inwards byup to 5 degrees or by up to 10 degrees relative to the radial direction.

For attachment to the transmission housing, according to an embodiment,the fastening ring comprises fastening regions in which fastening holesare provided. The fastening regions may be part of the fastening ring,or may be extensions that protrude from the fastening ring.

Advantageously, the attachment fixes the load cell both in the axialdirection and against a rotation in the circumferential direction,whereby the attachment can be released again mechanically. According toanother embodiment, this is achieved by a positive connection such as afaucet joint or a snap fit.

Connections that can not be easily untightened again, such as rivetjoints or welded joints are also possible. However, they are lesssuitable for maintenance purposes.

According to a further embodiment, the fastening ring comprisesrecesses, wherein the measuring regions are arranged in the recesses,wherein the recess can be formed in the radial and in the axialdirection. Thereby, it can be avoided, for example, that the measuringlugs directly abut the transmission housing or the thickness of thefastening ring can be adjusted independently of the thickness of themeasuring lugs.

According to a further embodiment, an angular extent of the measuringlugs and the slits delimiting the measuring lugs is approximately equalto an angular range of the axial support areas.

According to a further embodiment, an angular extent of the measuringlugs in the circumferential direction is less than or equal to 30degrees, so that the force can be measured in a defined direction.

The design of a load cell according to the present description isparticularly suitable for manufacturing the load cell integrally frommetal, for example by casting from a casting mold and mechanicalpost-processing steps.

In another aspect, the present description discloses a measuring devicefor determining a force acting on a crankshaft, in particular a pedalshaft. The measuring device comprises a crankshaft with a bearingarranged on the crankshaft, and a load cell according to one of theaforementioned embodiments. In this device, the receiving sleeve of theload cell is arranged on an outer ring of the bearing, wherein the axialsupport areas of the load cell are supported in the axial direction onthe outer ring of the bearing. Furthermore, an evaluation electronics isconnected to the strain sensors of the load cell.

Furthermore, the present description discloses a transmissionarrangement with the aforementioned measuring device. The transmissionarrangement comprises a transmission housing and a crankshaft, inparticular a pedal shaft. The crankshaft received or stored directly orindirectly in the transmission housing via a first drive side bearingand a second output side bearing. The bearings can be provided inparticular by rolling bearings.

Furthermore, the crankshaft may be passed through a hollow output shaft,wherein in particular for decoupling the pedal shaft from the hollowdrive shaft, a freewheel may be arranged between pedal shaft and thehollow drive shaft.

The first bearing is received in the transmission housing via the loadcell, wherein the load cell is received in the transmission housing orfastened to the transmission housing via the fastening ring. Thereceiving sleeve receives an outer ring of the first bearing, whereinthe axial support areas are supported on the outer ring of the firstbearing.

According to a further exemplary embodiment, the crankshaft of thetransmission arrangement comprises a larger diameter in the middle thanat its two ends. As a result, a first step and a second step of thecrankshaft are formed. An inner ring of the first bearing of themeasuring device abuts against the first step of the pedal shaft, and aninner ring of the second bearing abuts against the second step of thepedal shaft. As a result, an X-arrangement of a obliquely mountedbearing is formed and the load cell takes up a portion of the forces ofthe obliquely mounted bearing.

According to a further embodiment, the first bearing of the measuringdevice and the second bearing are each formed as a single-row angularcontact ball bearings. According to a further embodiment, the firstbearing of the measuring device and the second bearing each formed asobliquely mounted cylindrical roller bearing.

According to a further embodiment, the second bearing is supported via awave spring on the second step of the pedal shaft or on the housing oron a component connected to the housing.

According to a further embodiment, the second rolling bearing is furthersupported via a spacer disk on the second step of pedal shaft or on thehousing or on a component connected to the housing.

Furthermore, the present description discloses a transmissionarrangement which further comprises the following features:

a motor, a reduction gear connected to the motor, and a hollow outputshaft connected to the reduction gear. In this transmission arrangement,which is suitable for an electric bicycle, wherein the crankshaft isconfigured as pedal shaft, wherein the first bearing and the secondbearing are each configured as rolling bearings, wherein the pedal shaftis passed through the hollow output shaft, and wherein for decouplingthe pedal shaft from the hollow output shaft a freewheel is providedbetween the pedal shaft and the hollow output shaft.

Furthermore, the present description discloses an electrically drivenvehicle, in particular an electric bicycle, with the transmissionarrangement described above, wherein the motor is configured as anelectric motor, and wherein a battery of the electrically driven vehicleis connected to the electric motor.

In the following description, further details are given to describe theexemplary embodiments. It will be apparent to those skilled in the artthat the embodiments may be practiced without these details.

FIG. 21 shows a perspective view, viewed from the drive side, of a loadcell 47. The load cell 47 comprises four axial support lugs 91 on whichan outer ring of a ball bearing 45 is supported in the axial direction,and four measuring lugs 90 arranged between the axial support lugs 91,on each of which a strain gauge 92 is applied. For ease of positioningof the strain gauges 92, the surfaces of the measuring lugs 90 may berecessed.

The axial support lugs 91 comprise an outer portion 93 which rests on anouter ring of the ball bearing and an inner portion 94. The area inwhich the outer ring of the ball bearing rests can be seen in detail inthe cross-sectional view of FIG. 9.

The measuring lugs 90 and the axial support lugs 91 are each laterallyseparated by a milled radial slot 95. On the output side, the measuringlugs 90 transition into a receiving sleeve 96, which receives the outerring of the ball bearing 45. This receiving sleeve 96 is best seen inFIG. 26.

On the drive side, the lugs 90, 91 transition into an outer ring 97. Theregions of the outer ring 97 facing the measuring lugs 90 and the radialslots 95 each comprise, in the radial direction, a circumferential slot105 which has approximately half the radial extent of the measuring lugs90. A first radial slot 95, a circumferential slot 105, and a secondradial slot 95 together form a confining slot which extends in an angledU-shape and which separates the axial support lug 91 from the receivingsleeve 96 and from the adjacent measuring lugs.

The outer ring 97 comprises four fastening holes 98, with which the loadcell 47 can be secured to a transmission housing, not shown in FIG. 1,wherein a front side of the outer ring 97 rests against the transmissionhousing. An attachment of the load cell 47 to the transmission housingis shown by way of example in FIGS. 12 to 14.

FIG. 22 shows a plan view, seen from the drive side, of the load cell 47with the ball bearing 45 arranged in the load cell 47.

FIG. 23 shows a plan view, seen from the output side, of the load cell47 with the ball bearing 45 arranged in the load cell 47. As shown inFIG. 23, the measuring lugs 90 comprise portions 99 on the output sideend face of the outer ring 97 that are slightly recessed with respect tothe output side end face of the outer ring 97. As a result, thethickness of the measuring lugs can be reduced, resulting in a greaterdeformation.

FIG. 24 shows a cross-sectional view of the load cell 47 along thecross-sectional line A-A of FIG. 22, in which the recessed areas 99, thesleeve 96 and the outer ring 97 are visible.

FIG. 25 shows a side view of the load cell 47, in which slots 104 can beseen, which lie opposite the support lugs 91, and through which thesupport lugs 91 are separated from the sleeve 96.

FIG. 26 shows a perspective view of the load cell 47, viewed from theoutput side.

FIG. 27 shows a top view of the load cell 47, viewed from the outputside.

FIG. 28 shows a cross-sectional view of the load cell 47 taken along theintersection line A-A of FIG. 26.

FIG. 29 shows a cross-sectional view taken along the intersection lineB-B of FIG. 26.

FIG. 30 shows a harmonic pin ring transmission in which the load cell 47of FIG. 21 is incorporated and which will be described in detail below.In FIG. 30, the left side corresponds to a drive side and the right sidecorresponds to an output side of the harmonic pin ring gear 10.According to the usual arrangement of the drive on the right side, theviewing direction of FIG. 10 is directed in the direction of travel.

A stator 20 of a stator assembly of the harmonic pin ring gear 10 isdisposed in a motor housing 22. The stator 20 comprises three separatecoils 21 for connection to three phases of a three-phase inverter. Thethree coils of the stator 20 are connected to the three-phase invertervia three terminals 25, one of which is shown in FIG. 30.

The three-phase inverter is configured as a power electronics, which isarranged on a printed circuit board 23, wherein the printed circuitboard 23 is arranged in a cooling cover 24, which is mounted on themotor housing 22 on the drive side. The printed circuit board 23 isconfigured as an annular disk, which is located outside of cylindricalprotruding portions of the motor housing and the cooling cover, so thatthe motor electronics located thereon is sealed against oil and greaseof the transmission.

A pedal shaft 35 extending centrally through the motor housing 22 isstepped on the output side and comprises three steps whose diameterincreases from outside to inside and on each of which the shaft seal 50,the output side pedal shaft ball bearing 46 and the pedal shaftfreewheel 49 are arranged. The diameter of the pedal shaft 35 is alsostepped on the drive side and comprises two steps, wherein on the outerstep, the shaft seal and a sensor ring 68 are arranged, and on the innerstep, the ball bearing 45 is arranged.

An outer rotor shaft 26 equipped with permanent magnets is disposedradially inside the stator 20. This outer rotor shaft is also referredto as a “rotor package”. The outer rotor shaft 26 comprises on its innerside an elastic region which is plugged onto an inner rotor shaft 27. Onthe inner rotor shaft 27 an eccentrically arranged oval cam disk 28 isformed on the output side.

The inner rotor shaft 27 is supported in the motor housing 22 on thedrive side to the outside by a drive-side rotor ball bearing 29. Namely,an outer ring of the drive-side rotor ball bearing 29 is disposed in acylindrical recess of the motor housing 22.

Furthermore, the inner rotor shaft 27 is mounted on the output side inan output-side rotor ball bearing 30 radially outwardly in an inner gear7. A hollow shaft of the inner gear 7 is connected integrally with anannular portion of the inner gear 7 on the output side, which comprisesan external toothing 5.

The hollow shaft of the inner gear 7 is in turn mounted radiallyoutwardly via an inner gear ball bearing 31 on a housing cover 32 whichis screwed to the motor housing 22 by screws 33. The inner gear ballbearing 31 is offset from the output side rotor ball bearing 30 in theaxial direction to the output side and is offset in the radial directionto the outside. In addition, the inner gear ball bearing 31 overlapswith the output side rotor ball bearing 30 in the axial direction.

On the cam disk 28 of the inner rotor shaft 27, a flexible ball bearingor a thin section ball bearing 33 is clamped, in which an inner ring andan outer ring are deformable.

A pin retaining ring 103 with pins 101 bears on the flexible ballbearing 33, wherein the pins 101 are held in cylindrical recesses on theinside of a pin retaining ring. According to the embodiment of FIG. 30,the pins 101 are connected to each other. The pins 101 of the pin ring103 protrude the flexible ball bearing 33 and the pin retaining ring 103in axial direction on both sides. For the sake of simplicity, the pinring assembly comprising the pin retaining ring 103 and the pins 101will also be referred to below as pin ring 102.

The cam disk 28 and the flexible ball bearing 33 together form atransmitter arrangement, which converts a torque into a radial force.Instead of a flexible ball bearing with flexible inner and outer ring, awire-race bearing or a flexible ball bearing without outer ring, or adifferent kind of flexible rolling bearing can be used.

The housing cover 32 is screwed to the motor housing with fasteningscrews 34 on the output side of the motor housing. Furthermore, a driveside outer gear 8′ and an output side outer gear 8 are inserted into asupport ring 36 and screwed to the support ring 36 by the screws 34. Thesupport ring 36 is divided in the axial direction into two mutuallymirror-symmetrical halves, which together form a raceway 67 for the pinretaining ring 103.

The drive side outer gear 8′ and the output side outer gear 8 arearranged in the axial direction outside the cam disk 28 and the flexibleball bearing 33. In the radial direction, the drive side outer gear 8′is opposite to the areas of the pins 101 which protrude the pinretaining ring 103 on the drive side in the axial direction. In theradial direction, the output side outer gear 8 is opposite to the areasof the pins 101 which protrude the pin retaining ring 103 on the outputside in the axial direction.

On the drive side, a drive side spacer disk 37 is arranged in the motorhousing 22 such that it faces the drive side end faces of the pins 101in the axial direction. Similarly, on the output side, an output sidespacer disk 38 is arranged in the motor housing 22 such that it facesthe output side end surfaces of the pins 101 in the axial direction.

An output shaft 39 is disposed radially inside the hollow shaft of theinner gear 7, wherein a transmission freewheel 40 is disposed betweenthe hollow shaft of the inner gear 7 and the output shaft 39. The outputshaft 39 is mounted radially outwardly in an output side output ballbearing 41, which is inserted into a cylindrical recess or shoulder ofthe housing cover 32. The output side region of the output shaft 39protrudes the housing cover 32 in the axial direction. A chainringadapter 43 is mounted on the output shaft 39 and is held via a roundoutput nut 44 which is screwed into an internal thread of the outputshaft 39.

The motor housing 22 is sealed towards the transmission cover 32 by anO-ring 42 and towards the cooling cover 24 by a further O-ring 77.

FIG. 21 shows a detail of the pin-ring transmission of FIG. 19 in theregion of the load cell 47. As shown in FIG. 11, a sensor ring 68 isdisposed on the pedal shaft 35. In use, the sensor ring 68 may be usedto determine a position or a velocity of the pedal shaft 35.

For the sake of simplicity, it is assumed in the following that the loadcell is oriented such that the measuring lugs are respectively arrangedvertically and horizontally to the road surface, i.e. parallel to thepedaling movement of the rider and perpendicular thereto. Otherorientations are possible as well.

The load cell described above can be used in particular in atransmission of an electric bicycle to determine a measurement of theforce exerted on a pedal shaft and thus a required motor assistance. Theload cell does without moving parts and takes up little space,especially in the axial direction, which allows better use of theavailable space. In an electric bicycle, the space is limited and thusthe use of the load cell there may be particularly advantageous. Inparticular, the axial space is severely limited by the predeterminedoptimum distance of the pedal cranks.

During operation, a rider exerts forces on the pedals, in particular onthe down pedal. As a result, a downward force acts on the pedal shaft onthe side of the kicked pedal. Furthermore, a forward force acts on thepedal shaft on the side of the kicked pedal. Due to the lever arm on thebearing of the pedal shaft, a respective counteracting force acts on theopposite side of the pedal shaft.

When pedaling, the left pedal and the right pedal are alternatelykicked. Thus, with uniform cadence, a periodically alternating forceoccurs in the vertical direction and also in the horizontal direction.The amplitude of this force is correlated with the torque applied to thepedal shaft. When the measuring lugs are arranged vertically andhorizontally, respectively, the vertical force on the pedal shaft isdetected by strain gauges of the vertically aligned pair of measuringlugs and the horizontal force on the pedal shaft is detected by straingauges of the horizontally aligned pair of measuring lugs.

The electrical signal of a strain gauge is approximately equal andopposite to a signal of a radially opposite strain gauge. Thus, with anarray of radially opposed strain gauges, the measurement amplitude canbe doubled. This can be achieved by a subtractive superposition of thesignals, which can be done by analog electronics or after digitizationof the signal.

The signals of the strain gauges are transmitted via connecting lines toan evaluation electronics, which is arranged on a printed circuit boardwhich is fixed to the transmission housing. According to a simpleevaluation, an average torque is determined from one or more temporallyadjacent maximum deflections of the signals of the strain gauges inaccordance with a calibration curve which is stored in a permanentmemory of the evaluation electronics and an output signal is generatedtherefrom which correlates in a simple way with the torque exerted onthe pedal shaft, for example by proportional dependence.

According to a more elaborate evaluation, the further time developmentof the signals is also included in the electronic evaluation and anoutput signal is generated therefrom by means of previously stored datain the memory, such as calibration curves and parameters. Furthermore, acurrent angular position and/or a rotational speed of the pedal shaftmay be determined and included in the electronic evaluation.

The output signal can then be forwarded to a further part of theevaluation electronics, which determines a required motor assistance ofan electric motor of an electrically operated bicycle and generates acorresponding motor control signal. This motor control signal in turnmay depend on other parameters such as the angle of inclination of thevehicle, the current speed, a battery status or even certain drivingsituations that are derived from these parameters, such as driving overa curbstone, starting from a standstill, or starting on a slope.

On the one hand, a calibration can be carried out by directlycalculating a calibration curve and other calibration parameters fromthe component properties and storing them in the memory. On the otherhand, a calibration can also be carried out by attaching a furthersensor which measures the actual deformation of the pedal shaft and thusthe torque applied to the pedal shaft while at the same time a periodicpedal force is applied by a test device to the pedals of the pedalshaft.

Calibration parameters are determined from the correlation of theapplied pedal force and the actual torque. These calibration parameterscan then be stored in the memory of the evaluation electronics for allelectric bicycles of the same model. It is also possible to storecalibration parameters for different models in the same memory, whereinanother stored information indicates the currently used model of theelectric bicycle.

For the purpose of calibration, a measurement of the torque applied tothe pedal shaft can also be carried out without a sensor mounteddirectly on the pedal shaft, in that the torque at the output shaft orat downstream transmission elements is measured.

FIGS. 32 to 34 illustrate the absorption of radial and axial forces bytwo pedal shaft ball bearings, which are mounted on the pedal shaft inX-arrangement, and by a load cell 47. The viewing direction of FIGS. 21to 24 is in the direction of travel.

According to FIG. 32, the load cell 47 is fixed by screws which aretensile-loaded. However, the load cell can also be attached to anopposite side of a transmission housing, as shown in FIGS. 30 and 31. Inthis case, the forces directed axially outward are received directly bythe transmission housing.

FIG. 32 shows a transmission arrangement with a load cell 47 accordingto a first embodiment, wherein the load cell 47 is supported on adrive-side ball bearing 45 of an obliquely mounted bearing.

The obliquely mounted bearing comprises a pedal shaft 35, which has alarger diameter at the center than at its both ends, whereby adrive-side step 106 and an output side step 107 are formed. An innerring of a drive-side ball bearing 45 is supported in the axial directionon the drive-side step 106 and an inner ring of an output side ballbearing 46 is supported in the axial direction via a ring on the outputside step 107 of the pedal shaft 35.

This inwardly supported arrangement of the ball bearings 45, 46 on theshaft is also referred to as “X-arrangement”. The ball bearings 45, 46of the obliquely mounted bearing are configured as single-row angularcontact ball bearings, with the higher side of the respective inner ringpointing towards the center of the pedal shaft 35.

A support lug of the load cell 47 shown in cross-section is supported inthe axial direction on the output side step of the pedal shaft 35. Ameasuring lug 90 located behind it, to which a strain gauge 92 isattached, is supported in the radial direction on an outer ring of thedrive side ball bearing 45.

FIGS. 33 and 34 show a similar arrangement to FIG. 32, in which a radialportion of the measuring lugs 90 and the support lugs 91 lie in the sameplane. For clarification, the sectional plane of FIG. 13 extends througha support lug 91 and the sectional plane of FIG. 34 through a measuringlug 90 of the load cell 47. As shown in FIG. 34, an axial cross sectionof the measuring lug 90 in the region of the strain gauge 92 is narrowerthan the axial cross section of the support lug 91.

As a result, a larger deformation in the region of the strain gauge 92can be achieved. The thinning of the cross section can be achieved forexample by milling.

FIGS. 35 to 52 show further embodiments of a force measuring sensor fora pedal shaft. The force measuring sensor can be used to determineradial forces on a suspension of a pedal shaft. Indirectly, this can beused to determine a torque that a rider exerts on the pedal shaft. Thecorresponding device for determining the radial forces is for thisreason also referred to below as a torque measuring device.

FIG. 35 shows a further embodiment of a torque measuring device 110which is arranged in a bottom bracket bearing 109. The bottom bracketbearing 190 includes a pedal shaft 111 with fastening regions (not shownhere) for pedal cranks.

A cup-shaped sleeve 114 is disposed on the bottom bracket bearing 111between a first rolling bearing 112 and a second rolling bearing 113.The cup-shaped sleeve 114 is fixed at an end face 115 to the pedal shaft111. At an end opposite the end face, the sleeve 114 comprises a torquetransmitting portion 116.

A first strain gauge 117 is disposed between the sleeve 114 and thefirst rolling bearing 112 on the pedal shaft 111. A second strain gauge118 is disposed within the sleeve 114 on the pedal shaft 111, and athird strain gauge 119 is disposed on an outer surface 120 of the sleeve114.

The strain gauges 117, 118, 19 are each separately electricallyconnected to a slip ring 121, which is arranged on the outside of thepedal shaft 111. A torque flow from the pedal shaft 111 via the sleeve114 is indicated in FIG. 15 by arrows.

FIG. 36 shows a second embodiment of a torque measuring device 110′. Incontrast to the first embodiment, slip rings 121 to which a strain gauge119 is connected on the outer surface 120 of the sleeve 114 are disposedon the outer surface 120 of the sleeve 114.

FIG. 37 shows a third embodiment of a torque measuring device 110″. Incontrast to the previous embodiments of FIGS. 35 and 36, strain gauges117, 118 and 119 are connected to a transmitter 122.

FIG. 38 shows a fourth embodiment of a torque measuring device 110′″.According to this embodiment, a strain gauge 124 is disposed on an outerring of the right ball bearing 112. Alternatively, a strain gauge canalso be arranged on an intermediate ring.

FIGS. 39 and 40 show another embodiment of a torque measuring device 130in which a deformation on a transmission housing is measured.

In the torque measuring device 130, a first pair of radially opposedstrain gauges 131, 132 is arranged on a housing 135 and a second pair ofradially opposed strain gauges 133, 134 is arranged on the housing 135and offset by 90 degrees from the first pair of strain gauges 131, 132.

Furthermore, a printed circuit board 136 is arranged on the housing 35,on which an evaluation logic for the signals of the strain gauges 31,32, 33, 34 is provided. By displacing the individual strain gauges by 90degrees, the arrangement can be installed in any orientation or angularposition. This does not exclude that there are preferential orientationsthat are better suited than others.

FIG. 40 shows another embodiment of a torque measuring device 130′ whichis similar to the embodiment of FIG. 39. Unlike the embodiment of FIG.39, the second pair of strain gauges is offset only slightly from thefirst pair of strain gauges, for example by approximately 10 degrees.

FIG. 41 shows a first sequence of measurement values of a torque sensoraccording to FIG. 39.

FIG. 42 shows a further sequence of measurement values of a torquesensor according to FIG. 39. In this case, a value range of themeasurement signal in millivolts is plotted on the right-hand axis and avalue range of a force derived from the measurement signal inNewton-meters is plotted on the left-hand axis.

FIG. 43 shows a cross-section of another HCD transmission 10′ with aload cell 47 similar to the transmission 10 of FIG. 30.

Unlike the transmission of FIG. 40, a magnetic transmitter ring 137 isintegrated in a side cover of the ball bearing 45 connected to the innerring of the ball bearing 45. A rotational speed sensor 138 lyingopposite the transmitter ring 137 in the axial direction registerschanges in the magnetic field caused by the rotation of the transmitterring 137. In particular, the transmitter ring 137 may be magnetized suchthat north and south poles alternate, so that the rotational movement ofthe transmitter ring 137 generates a periodically variable field at thelocation of the rotational speed sensor 138.

Furthermore, the pedal shaft 35 is configured as a hollow shaft, whichcomprises an enlarged circumference only in the areas in the vicinity ofthe steps. The outer rotor shaft 26 is screwed onto a thread 140 of theinner rotor shaft 27. The inner rotor shaft 27 comprises a drive sidestep 141, on which the ball bearing 29 is supported. The cam disk 28comprises an annular protrusion 139, by which the cam disk 28 issupported on the drive side in the axial direction on the outer rotorshaft 26.

Thus, the axial position of the outer rotor shaft 26 is supported in theoutput side direction by the thread 140 of the inner rotor shaft 27 andin the output side direction by the protrusion 139 of the cam disk 28. Afirst drive side power flow runs from the cam disk 28 via the outerrotor shaft 26 and the thread 140 into an output side region of thepedal shaft 35. A second drive side power flow runs from a drive sidestep 141 of the inner rotor shaft 27 via the ball bearing 29 and thewave spring 61 into the transmission housing 22.

Unlike the transmission of FIG. 30, a pedal shaft freewheel 49′ isconfigured as a clamp roller freewheel in the embodiment of FIG. 43,similar to the motor freewheel 40. A clamp roller retaining ring of theclamp roller freewheel 49′ is placed on a step of the pedal shaft 35 andsupported by an O-ring 142.

FIG. 44 shows an enlarged region of FIG. 43 in the region of the loadcell 47.

Although the above description contains many details, these should notbe construed to limit the scope of the embodiments, but only as anillustration of anticipated embodiments. In particular, theabove-mentioned advantages of the embodiments should not be construed tolimit the scope of the embodiments, but only as an illustration ofpossible effects when the described embodiments are put into practice.Accordingly, the scope of the embodiments should be determined by theclaims and their equivalents rather than by the examples described.

FIGS. 45 to 60 show a pedal shaft assembly and a harmonic pin ringtransmission with such a pedal shaft assembly.

The subject matter of the present description will be further withreference to the following figures

The following description mentions details for describing theembodiments of the present description, such as the shape and number ofparts of the freewheels. It should be apparent to those skilled in eachcase, if the embodiments can be implemented without these details inpractice.

FIG. 45 shows an exploded view of a pedal shaft assembly 80, in whichfrom a drive side to an output side, an inner gear 7, coil springs 66 ofan external or transmission freewheel 40, cylindrical clamp rollers 64of the transmission freewheel 40, a pedal shaft 35, coil springs 72 ofan inner or pedal shaft freewheel 49, pawls 73 of the pedal shaftfreewheel 49, an output shaft 39 and a freewheel cage 65 of thetransmission freewheel 49 are shown.

The output side of the pedal shaft assembly 80 is here to be understoodas the side on which a receiving region 220 of the output shaft 39 foran output is located. Accordingly, the drive side is the side oppositethe output side.

The radius of the pedal shaft 35 is stepped such that two steps 222, 223are formed on the drive side and three steps 224, 225, 226 are formed onthe output side. The steps 222, 223, 224, 225, 226, which are best seenin FIG. 3, form receiving areas for further transmission elements, whichare not shown in FIG. 45. Furthermore, a star arrangement 227 withspikes 71 is formed on the output side on the third step.

The transmission freewheel 40 is also referred to as “motor freewheel”,which is useful for differentiation when the pedal shaft 35 is connectedto the output shaft 39 via an intermediate gear such as a planetarygear.

The outside of the output shaft 39 comprises a stair-shaped rollingregion 228 for the clamp rollers 64 and recesses for the coil springs 66on the drive side. The inside of the output shaft 39 or the hollowoutput shaft 39 comprises a stair-shaped stop portion 229 for the pawls73 radially opposite to the stair-shaped rolling region 128

Furthermore, the output shaft 39 comprises a receiving region 220 for anoutput means (not shown in FIG. 45) on the output side wherein theoutput means may be in particular a chainring adapter, and comprises aninternal thread for fixing a transmission cover (not shown in FIG. 45)radially opposite to the receiving region for the output means.

FIG. 46 shows a side view of the pedal shaft assembly 80 from the outputside, in which a side cover of the freewheels is removed for ease ofillustration of the transmission freewheel 40 and the pedal shaftfreewheel 49. In this view, seen from the inside outward, the outputside receiving region 75 for a pedal crank (not shown in FIG. 46), thestar arrangement 227 with the pawls 73 and the coil springs 66, thestair-shaped stop portion 229 of the pedal shaft freewheel 49, thereceiving region for the output means, the freewheel cage 65 of thetransmission freewheel 40, the pinch rollers 64 of the transmissionfreewheel, the outer ring 65 of the transmission freewheel 40, and theinner gear 7 of a harmonic pin ring gear (not shown in FIG. 46) areshown.

FIG. 47 shows a cross-sectional view of the pedal shaft assembly 80 ofFIG. 45 taken along the cross-sectional line shown in FIG. 46. As shownin FIG. 3, an inner radius of the output shaft 39 configured as a hollowshaft is stepped and forms three steps 230, 231, 232. The outermost stepof the output shaft 39 comprises the internal thread 233. The secondoutermost step of the output shaft 39 and the opposite second outermoststep of the pedal shaft 35 serve to receive a ball bearing, not shown inFIG. 47.

On the inner gear 7, a hollow shaft 234 is formed on the output side,which at the same time forms an outer ring of the transmission freewheel40. Furthermore, the inner gear 7 comprises a disk-shaped region 235with an external toothing 5 on the drive side.

On the drive side, an annular thickening is formed on a side edge of theoutput shaft 39, with which the drive shaft 39 is inserted into thehollow shaft 134 of the inner gear 7. A radius of the outer side of theoutput shaft 39 is also stepped, wherein the radially outermost step isadapted in the radial direction for receiving the freewheel cage 65. Thefurther steps of the outer side of the output shaft 39 are configuredfor receiving a ball bearing (not shown here) including a shaft seal andfor receiving an output means (not shown here).

FIG. 48 shows a perspective view, seen from the output side, of thepedal shaft assembly 80 of FIG. 45, in which a side cover is removed andin which the stop region of the pedal shaft freewheel 49, the freewheelcage 65 and the clamp rollers 64 are partially visible.

FIG. 49 shows a further perspective view of the transmission assembly ofFIG. 45 viewed from the drive side, and in which in particular the innerstair-shaped portion 229, the pawls 73, the coil springs 72 and theouter stair-shaped portion 228 of the pedal shaft freewheel 49 and theclamp rollers 64, the freewheel cage 65 and the coil springs 66 of thetransmission freewheel 40 can be seen.

FIG. 50 shows a side view of the pedal shaft 35 of the pedal shaftassembly 80 of FIG. 45, particularly showing the crankshaft receivingareas 74, 75, the stair-shape of the circumference of the pedal shaft35, and the star assembly 227.

As shown in FIGS. 50 and 52, the spikes 71 of the star assembly 227, ofwhich four spikes 71 are visible in FIG. 50, each comprise a rollingregion 243, an end portion 244, and a pawl receiving region 245. Thespikes 71 each comprise a bore 236 for receiving the coil spring 72 (notshown in FIG. 50), wherein the bore 136 is aligned perpendicular to therolling region 233. The bore 236 is slightly offset from a center of therolling region 243 toward the pawl receiving region 245.

FIG. 51 shows a side view of the output shaft 35, in which in particularthe rolling region 233 with the receiving region for the coil spring 72(not shown in FIG. 7), the step 228 of the outer circumference and thereceiving region 220 for the output means are shown.

52 shows a side view of the pedal shaft 35 seen along the central axisof the pedal shaft 35 from the drive side, in which in particular thereceiving region 75 for a pedal crank, the output-side stair-shape ofthe outside diameter of the pedal shaft 35 and the star arrangement 227are shown. The end portions 234 of the spikes 71 comprise a chamfer 237on one side.

During operation, the coil springs 72 press the pawls 73 outwardlyagainst the stair-shaped portion 229 of the inside of the output shaft39, so that the tip of the pawl 73 engages an opposing step when thepedal shaft 35 moves in a drive direction which is given by theorientation of the pawls 73 and the orientation of the steps, fasterthan the drive shaft 39.

If the pedal shaft 35 moves slower in the drive direction than the driveshaft 39 or even in the opposite direction to the drive shaft 39, thepawls 73 slide along the steps and are pressed inwards against thespring force of the coil springs 72 until the next step is reached. Atthe transition of two steps, the pawl 73 jumps to the outside by theaction of the coil spring 72 and causes in this way a characteristicclick.

The click of the pedal shaft freewheel 49 can fulfill a warning functionin that passers-by are made aware of an electric bicycle that drivesrelatively quietly, unlike a moped or scooter. Furthermore, it allows afunction control due to the noise and can meet the expectation ofcustomers who are used to an idling noise.

FIG. 53 shows an output side view of the output shaft 39. A drivedirection is predetermined by the direction of the step heels and pointsin the clockwise direction in the view of FIG. 53. In this case, lessinclined step portions of the inner 229 form the step heels and the moreinclined step heels form step stops.

The steps are aligned so that the drive direction corresponds to thedirection of travel when the output is provided in the conventional wayin the direction of travel on the right. However, the output can also beprovided inversely in the direction of travel left. This is particularlypossible in the case of three- or four-wheelers. In this case, thedirection of the steps must be reversed with respect to the arrangementof FIG. 9 in order to allow a drive in the direction of travel.

In the case of three- or four-wheelers, a switchable freewheel may alsobe expedient, or an additional drive connection such as a switchableclutch which allows reverse drive of the motor, for example to drivebackwards onto a ramp, in particular if the vehicle is equipped to carryloads.

FIG. 54 shows the plan view corresponding to FIG. 53 on the output shaft35 from the drive side.

FIG. 11 shows a side view of the freewheel cage 65, wherein the viewingdirection is perpendicular to the central axis of the pedal shaft 35.FIG. 12 shows a side view of the transmission freewheel 40, in which theclamp rollers 64 and the freewheel cage 65 are partially shown.

The freewheel cage 65 comprises webs 150 uniformly distributed on thecircumference and two opposite receiving regions 251, 252 for the coilsprings 66, which are aligned with the recesses of the output shaft 39.

FIG. 57 shows a side view of a pedal shaft assembly 80 which comprisesan inner portion of the freewheel assembly 81 shown in FIG. 1 and asensor assembly not shown in FIG. 1.

FIG. 58 is a cross-sectional view taken along the cross-sectional lineD-D shown in FIG. 13.

FIG. 59 is a cross-sectional view similar to the cross-sectional view ofFIG. 14, but in which the viewing direction is directed from the driveside. In this case, the hollow shaft portion 235 of the inner gear 7,which is not shown in FIG. 58, is additionally shown, which forms theouter ring of the transmission freewheel 40.

The pawls 73 of the pedal shaft freewheel 49 comprise on one side acylindrical hinge portion 246 which engages in a mating round hingeportion 247 of the stair-shaped portion. Furthermore, the pawls comprisea plate-shaped portion 248, which merges at one end into the hingeportion 247 and is chamfered to a pointed edge 249 at an end oppositethereto.

During assembly or maintenance of the pedal shaft freewheel 49, thepawls 73 can be easily inserted from the output side in the round hingeportion 247 formed from the pedal shaft 35 without the need for furthercomponents such as an axle to form the required hinge.

The two springs 66 of the transmission freewheel 40 push, via thefreewheel cage 65, all the clamp rollers 64 almost in the end ortraction position shown in FIG. 59. As a result, the clamp rollers 64come into contact with the outer ring formed by the hollow shaft portion234 of the inner gear 7 and can be moved to the end position by relativemovement of the outer ring. In this case, the individual webs of thecage 65 may be formed elastically, and thereby compensate for anon-uniform contact pressure on the individual clamp rollers 64.

However, if the outer ring of the transmission freewheel 40 moves slowerthan the output shaft 39 in the drive direction, the clamp rollers 64roll towards the bottom of the steps and are thereby lifted from theouter ring, so that the frictional connection with the inner gear 7 isremoved.

FIG. 60 shows a cross section of a harmonic pin ring gear 10′, in whicha pedal shaft freewheel 49′ is configured as a clamp roller freewheel.Accordingly, in this embodiment, there is no step-shaped engagementregion for pawls on the inside of the output shaft.

A clamp roller retaining ring of the clamp roller freewheel 49′ isplaced on a step of the pedal shaft 35 and is supported by an O-ring262.

FIGS. 61 to 66 show a geared motor with a harmonic pin ring transmissionhaving an eccentric disk.

FIG. 61 shows a cross-sectional view of a harmonic pin ring transmission10. The cross-sectional plane A-A of FIG. 61 is indicated in FIG. 62. InFIG. 61, the left side corresponds to a drive side and the right sidecorresponds to an output side of the harmonic pin ring gear 10.

A stator 20 of a stator assembly of the harmonic pin-ring transmission10 is disposed in a motor housing 22. The stator 20 comprises a statorcoil 21 (not shown in FIG. 61) for connection to a power supply.

An outer rotor shaft 26 equipped with permanent magnets is disposedradially inside the stator 20 on an inner rotor shaft 27. On the innerrotor shaft 27, an eccentrically arranged circular eccentric disc 28′ oran eccentric circular disc 28′ is formed on the output side, which isshown in more detail in the exploded view of FIG. 64.

On the drive side, the inner rotor shaft 27 is supported to the outsideby a drive side rotor ball bearing 29 in the motor housing 22. An outerring of the drive side rotor ball bearing 29 is arranged in acylindrical recess of the motor housing 22. On the drive side, a sensorring 68 is disposed on the inner rotor shaft 27 axially adjacent to theouter rotor shaft 26, which is opposed to a Hall sensor 353 disposed inthe motor housing 22.

On the output side, the inner rotor shaft 27 is supported in an outputside rotor ball bearing 30 radially outward in an inner gear 7. A hollowshaft of the inner gear 7 is connected integrally with an annularportion of the inner gear 7 on the output side, which comprises anexternal toothing 5.

On the output side, a housing cover 32 is screwed by screws 34 to themotor housing 22. On the opposite drive side, an inner housing 346 isscrewed to the motor housing 22 by screws 347. The inner housingcomprises a support cylinder 348.

The hollow shaft 52 of the inner gear 7 is in turn mounted radiallyoutwardly via an inner gear ball bearing 31 on the housing cover 32. Theinner gear ball bearing 31 is offset from the output side rotor ballbearing 30 in the axial direction to the output side and is offset inthe radial direction relative to the output side rotor ball bearing 30to the outside. In addition, the inner gear ball bearing 31 overlapswith the output side rotor ball bearing 30 in the axial direction.

On the eccentric circular disk 28 of the inner rotor shaft 27, atransmitter ball bearing 33 is arranged. The eccentric disk 28′ and thetransmitter ball bearing 33 together form a transmitter arrangementwhich converts a torque into a radial force. On the transmitter ballbearing 33 a pin retaining ring 103 with pins 101 is mounted, whereinthe pins 101 are held in cylindrical recesses on the inside of the pinretaining ring 103. The pins 101 of the pin ring 103 protrude the ballbearing 33 and the pin retaining ring 103 in the axial direction on bothsides. For the sake of simplicity, the pin-ring arrangement of pinretaining ring 103 and pins 101 will also be referred to below as pinring 102.

A drive side outer gear 8′ and an output side outer gear 8 are insertedinto webs 349 of the motor housing 22. The webs 49 of the motor housing22 are shown in the exploded view of FIG. 63. Between the motor housing22 and the housing cover 32, an O-ring 42 is disposed on the output sideradially outward. Another O-ring 343 is disposed on the drive sidebetween a step of the inner rotor shaft 27 and the drive side rotor ballbearing 29.

The drive side outer gear 8′ and the output side outer gear 8 arearranged in the axial direction outside the eccentric disk 28′ and thetransmitter ball bearing 33. In the radial direction, the drive sideouter gear 8′ is opposite to the areas of the pins 101 which protrudethe pin retaining ring 103 on the drive side in the axial direction. Inthe radial direction, the output side outer gear 8 is opposite to theareas of the pins 101 which protrude the pin retaining ring 103 on theoutput side in the axial direction.

On the drive side, a drive side spacer disk 37 is arranged in the motorhousing 22 such that it faces the drive side end faces of the pins 101in the axial direction. Similarly, on the output side, an output sidespacer disk 38 is arranged in the motor housing 22 such that it facesthe output side end surfaces of the pins 101 in the axial direction. Thehollow shaft 352 of the inner gear 7 comprises screw holes (not shown inFIG. 61) for connecting an output.

Between the inner cylinder 348 of the inner housing 46 and the housingcover 32, an inner shaft seal 50 is inserted. Furthermore, an outershaft seal 51 is disposed opposite the output side pedal shaft ballbearing 46 between the hollow shaft of the inner gear 7 and the housingcover 32.

In operation, an input torque is transmitted via the stator 20 byelectromagnetic force action to the outer rotor shaft 26 and from thereto the inner rotor shaft 27, which is converted by the eccentric disk28′ and the flexible ball bearing 33 into a radial force. This radialforce is converted at the tooth flanks of the internal toothings 6, 6′of the outer gears 8, 8′ and the external toothing 5 of the inner gear 7into an output torque, wherein the inner gear 7 is driven, and the outergears 8, 8′ are fixed to the housing. The output torque is larger thanthe input torque by the reduction gear ratio.

The inner toothing, which is formed by the external toothing 5 of theinner gear 7, lies opposite the internal toothing 6 of the output sideouter gear 8, and thereby provides the output torque, in particular bythose pins 101 which abut on both the external toothing 5 and theinternal toothings 6, 6′.

FIG. 63 shows a drive side part of an exploded view of the transmissionof FIG. 61, in which, viewed from the drive side to the output side, theinner housing 346 with the inner cylinder 348, the drive side rotor ballbearing 29, the round-wire wave spring 44, the motor housing 22 with thedrive side spacer disk 37, the sensor ring 68 and the rotor shaft 26 areshown.

FIG. 64 shows an output side part of an exploded view of thetransmission of FIG. 61, in which, seen from the drive side to theoutput side, the rotor shaft 26 with the eccentric disc 28′, the secondouter gear 8′, the pin ring 102 with the pins 101 and the pin retainingring 103, the transmitter ball bearing 34, the output side rotor ballbearing 30, the first outer gear 8, the shaft seal 50, the inner gearwith the inner gear hollow shaft, the driven side spacer disk 38, theinner gear ball bearing 31, the transmission cover 32 and the shaft seal51 are shown.

The outer gears 8, 8′ each comprise grooves 354 which are radiallydirected radially inwardly from the circumference of the respectiveouter gear 8, 8′ and are distributed at regular intervals over thecircumference of the respective outer gear 8, 8′. The motor housing 2comprises circumferentially spaced journals 355 corresponding to thegrooves 354.

As shown in FIG. 61, in the assembled state, the screws 34 extendthrough the transmission cover 32, the first outer gear 8, the secondouter gear 8′, and the support ring 36 into the motor housing 22.

FIGS. 67 to 76 show a harmonic pin ring transmission with a crank gear.The crank gear is a planetary gear, which is arranged on the pedal shaftand transmits the cadence of a rider into the fast. The planetary gearcan also be formed switchable, in particular if the planetary gearincludes further gear stages, which are not shown in FIGS. 67 to 76, asit is the case for example in a Ravigneaux or Lepelletier planetarygear.

In the simplest case, the shifting allows a choice between a 1 to 1transmission and a speed-up transmission. The shifting of the crank gearcan be done manually via an operating element or automatically. In thecase of an automatic shifting, the shifting can take place on the basisof measured values of a torque sensor.

Compared to the embodiments of FIGS. 1 to 40, the bottom bracket bearingsensor unit is replaced by a crank gear. By transmitting the planetarygear into the fast, a secondary transmission to the rear wheel can bemade smaller. This in turn makes it possible to rotate the output shaftfaster, so that the motor and/or the reduction gear can be made smalleror deliver more power with the same dimension.

Furthermore, the planetary gear can be used to measure the torque on thepedal shaft or pedal crank by measuring the support force of an elementfixed to the housing of the planetary gear, for example by glued straingauges.

If the sun gear is fixed to the housing and the ring gear is driven, asin the embodiment of FIGS. 72 to 76, for example a speed-up transmissionof 1.59:1 can be provided, which in turn allows reducing the outputmotor power by a factor of 0.63. If the ring gear is fixed to thehousing and the sun gear is driven, as in the embodiment of FIGS. 67 to71, for example a speed-up transmission of 3:1 can be provided. Thus,the output torque of the motor can even be reduced by a factor of 0.33.

The transmission of the planetary gear can be used for example toincrease the power density, which, however, can also increase the motorspeed and thus the motor noise. Furthermore, the planetary gear can beused as a basis for a motor with integrated change speed gear. In thiscase, the change speed gear can be dimensioned correspondingly smalleror the necessary space for the change speed gear can be made availabledue to an increased motor power density.

FIGS. 67 to 71 show a harmonic pin ring transmission with a planetarygear arranged on a pedal shaft, in which a drive takes place via aplanet carrier which is positively connected to the pedal shaft, and inwhich an output takes place via a sun gear.

FIGS. 72 to 76 show a harmonic pin ring transmission with a planetarygear arranged on a pedal shaft, in which a drive takes place via aplanet carrier rotatably mounted on the pedal shaft, and in which anoutput takes place via a ring gear rotatably mounted in the transmissionhousing.

The transmissions are similar to the harmonic pin ring transmission ofFIG. 11, which comprises an eccentric disk. However, the crank gears ofFIGS. 67 to 76 may be combined with other reduction gears as well. Inparticular, the crank gears of FIGS. 67 to 76 may be combined with aharmonic pin ring transmission that is similar to the transmission ofFIG. 1 and comprises an oval shaped cam disk and a flexible ballbearing.

For the sake of clarity, not all of the components already shown in FIG.1 or in FIG. 11 have been provided with reference numerals in FIGS. 69to 76 again.

FIG. 68 shows a perspective view of a planetary gear assembly 400 ofFIG. 67, in which a ring gear 401, a mounting sleeve 402 of the ringgear 401 and an output shaft 39 can be seen. FIG. 69 shows a side viewof the planetary gear assembly 400 of FIG. 68.

FIG. 70 shows a cross-section of the planetary gear assembly 400 takenalong intersection line C-C of FIG. 69. A sun gear 403 is configured asa hollow shaft which is separated by a gap 404 from the pedal shaft 35so that the sun gear 403 can rotate relative to the pedal shaft 35.Planet gears 405 are respectively disposed on planetary axes 406,wherein a rolling bearing or a sliding bearing, which is formed by asliding film, is arranged between the planet gears 405.

FIG. 71 shows a cross-section of the planetary gear assembly 400 takenalong intersection line B-B of FIG. 69. In this cross-sectional view isfurther shown that the planetary axes 406 are disposed in a planetcarrier 407, which is positively connected to the pedal shaft 35.

The sun gear 403 is integrally connected to another hollow shaft 410,which comprises a slightly larger diameter than the sun gear 403, whichis dimensioned such that a rolling bearing 408, which comprises twoneedle bearings in the embodiment of FIG. 71, can be arranged betweenthe hollow shaft 410 and the pedal shaft 35. On the second hollow shaft410, a pedal shaft freewheel 49 is arranged.

A torque flow from the pedal shaft 35 to the output shaft is indicatedin FIG. 71 by arrows. As shown in FIG. 67, the fastening flange 402 ofthe ring gear is fixed to the transmission case 22. The stationary ringgear absorbs the opposing forces caused by the rider's pedaling motion.

According to a further embodiment, deformation sensors such as straingauges, which are connected to an evaluation electronics, are mounted onthe ring gear 401. For a more accurate measurement, a portion of thering gear, in which the deformation sensors are mounted, may be madethinner.

FIG. 72 shows a harmonic pin ring transmission with a planetary geararranged on a pedal shaft 35 in which an output takes place via a ringgear.

FIG. 73 is a perspective view of the planetary gear assembly 400′ ofFIG. 67. FIG. 74 shows a side view of the planetary gear assembly 400′of FIG. 68.

FIG. 75 shows a cross section along the intersection line C-C of FIG.74. A sun gear 403′ is separated from the pedal shaft by a gap 404. Asshown in FIG. 72, the sun gear 403′ is connected to the transmissioncase 22 via a fastening region. In particular, this fastening region maybe configured as a load cell. The planet gears 405 are rotatably mountedon planetary axes 406 disposed in a planet carrier 405.

FIG. 76 shows a cross section taken along intersection line B-B of FIG.74. As shown in FIG. 76, a portion of the planet carrier 405 is formedas a hollow shaft 409 which is disposed on the pedal shaft 35 via apedal shaft freewheel 49′.

According to a further embodiment, which is not shown here, a hollowshaft of a planetary gear arranged on the pedal shaft is fixed to thehousing on the drive side, a hollow shaft of the planetary gear isfixedly connected to the pedal shaft, and a sun gear is coupled to anoutput shaft via a pedal shaft freewheel. To this end, an attachment ofthe hollow shaft may be guided around the planetary gear from a driveside. Here and in the other crank gears the pedal shaft may alsogenerally be a crankshaft, such as a crankshaft of an internalcombustion engine, or a drive shaft of a drive. Just like the two crankgears mentioned above, this crank gear also provides a speed-uptransmission.

In addition, a reversal of rotation direction arises, which may befavorable in cases where, unlike an electric bicycle without reversegear, a direction of rotation of a load is set in opposite direction tothe direction of rotation of the crankshaft or drive shaft of the crankgear, so that in these cases no further reversal of rotation directionis necessary.

FIGS. 77-81 show a cycloidal gear according to the present descriptionand a motor gear unit with the cycloidal gear.

FIG. 77 is a cross-sectional view of a motor gear unit for an electricbicycle in which a cycloidal gear is used as a reduction gear.

Components which have already been explained in the description of theprevious figures, in particular FIGS. 1 to 3 and 10 to 14, will not bedescribed again here. Like the transmissions shown in FIGS. 1 and 10,the cycloidal gear comprises a three-bearing arrangement in which anoutput element is formed in one piece with a hollow output shaft and issupported on two diagonally opposite bearings inwardly on a housingcover and to the outside on an inner hollow shaft. Thus, only threebearings are required to support the rotor shaft, the output element andthe hollow output shaft.

In the three-bearing arrangement, the outer bearings of the rotor shaftcan be further apart. This is the case in particular in a motor gearunit for an electric bicycle, in which a limited axial space is utilizedbetter by using fewer bearings. Thus, a leverage effect of an outertilting moment on the bearings and on the rotor shaft is smaller, andthe rotor shaft can be constructed as a thin cylinder. This is the casewith many such engine gear units.

In the cycloidal gear of FIG. 77, the output element on which the hollowoutput shaft is formed, is formed by an output pulley with carrier pinsand carrier rollers arranged thereon. In the motor gear unit shown inFIG. 1 and FIG. 10, it is formed by an output inner gear on which ahollow output shaft is formed.

An inner rotor shaft 27′ of the cycloidal gear is formed of a driveshaft 426 and a rotor shaft 427 mounted thereon. The drive shaft 426comprises drive side circular eccentric disk 428, an output sidecircular eccentric disk 429 and a centered circular disk 430, which areformed on the drive shaft 426 and which are arranged next to oneanother. The output side eccentric disk 429 is offset from the driveside eccentric disk 428 by 180 degrees.

On the drive side eccentric disk 428, a first ball bearing 423 isarranged, on which a drive side inner gear 433 is mounted. An outputside inner gear 434 is mounted on the output side eccentric disk 429 viaa second ball bearing 424. The drive side inner gear 433 and the outputside inner gear 434 are identical in construction and each comprise anexternal toothing 435, 436, which engages in each case in an oppositeinternal toothing 437 of an outer gear 439 stationarily connected to thehousing.

An output disk 440 is mounted on a third ball bearing 425 via a ring 441disposed on the centered circular disk 430. The output pulley 440comprises carrier pins 442 which are arranged at regular intervals onthe output pulley 441 and which engage in circular openings 444 of thetwo inner gears 433, 434. The carrier pins 442 are provided with rollers443 which are rotatably mounted on the carrier pins 442.

The drive shaft 426 of the inner rotor shaft 27′ acts as a hollow driveshaft 426 of the cycloidal reduction gear.

On the rotor shaft 427 of the inner rotor shaft 27′, a pressure disk 451and a thread 450 opposite the pressure plate are formed. As can be seenin the cross-sectional drawing of FIG. 80, a pressure ring 452 isscrewed onto the thread, so that the rotor packet 26 or the outer rotorshaft 26 is clamped between the pressure ring and the retaining ring.This corresponds to the arrangement of FIG. 30 and differs from thearrangement of FIG. 1, in which the rotor packet is held between aspacer arranged on the inner rotor shaft and a shoulder of the innerrotor shaft.

The drive side inner gear 433 and the output side inner gear 434 arealso referred to as “cam disks”. Instead of the internal toothing 437 ofthe outer gear 439, a stationary pin ring can also be provided, which isindicated in FIG. 79. This stationary pin ring may have rollers, so thatfriction and shear forces are reduced and a rolling movement ispossible.

The exploded view of FIG. 78 shows, viewed from left to right, the outergear 439 with the internal toothing 437, the output side inner gear 433,the first ball bearing 423, the inner rotor shaft 27′ on which the driveside eccentric disk 428, the output side eccentric disk 429 and thecentered circular disk 430 are arranged, the second ball bearing 424,the output side inner gear 434, a spacer 445, the ring 441, the thirdball bearing 425, the carrier rollers 443 and the carrier pins 442arranged on the output pulley 440, and the output pulley 440.

Furthermore, FIG. 78 shows a ridge 446 which is provided between thedrive side 428 and the output side eccentric disk 429.

The side view of FIG. 79, seen from inside to outside, shows the ring441, the third ball bearing 425, the output pulley 440, the output sideinner gear 434 with the external toothing 436, the external toothing 435of the drive side inner gear 433, and the internal toothing 437 of theouter gear 439. Stationary pins 447 are indicated in two positions,which may be provided instead of the internal toothing of the outer gear439.

The cross-sectional view of FIG. 80 shows a cross section through thecycloidal gear of the motor gear unit of FIG. 77, in which the innerrotor shaft 27′ with the drive side eccentric disk 428 and the driveside inner gear 433 mounted thereon, the output side eccentric disk 429and the output side inner gear 434 mounted thereon, the centeredcircular disk 430 and the output disk 440 mounted thereon with thecarrier pins 442 and the carrier rollers 443, and the outer gear 439with the internal toothing 437 are shown.

In the lower part of FIG. 80, a fastening opening 448 of the outer gear439 is shown. As shown in the cross section of FIG. 77, the outer gear439 is fixed to the transmission case by fastening bolts that are passedthrough the fastening openings 448.

FIG. 81 shows a side view of a detail of the cycloidal gear of FIG. 77,in which also hidden components are visible. In the view of FIG. 81 itcan be seen that all the tooth heads of the two inner gears are inengagement or in contact with the internal toothing of the stationaryouter gear. Furthermore, FIG. 81 shows at the same time an alternativeembodiment in which, instead of an internal toothing, the outer gearcomprises an arrangement of stationary pins or bolts with rollersarranged thereon.

In the following, an assembling of the cycloidal drive will be describedby way of example with reference to the preceding FIGS. 77-81.

The elements that are on the drive side of the rotor shaft 427 in themotor housing 22 are inserted into the motor housing 22 or attached tothe motor housing 22. Among other things, the drive side rotor ballbearing 29, the load cell 47, the drive side pedal shaft ball bearing 45and a stator assembly of the motor are inserted into the motor housing22. The outer gear 439 is screwed onto the motor housing 22.

The rotor shaft 427 is attached to the output shaft 426 on the driveside. The drive side inner gear 433 and the first ball bearing 423 areplaced on the drive side eccentric disk 428 from the drive side. Therotor packet is placed on the inner rotor shaft 27′ from the drive sideand the pressure ring 452 is screwed to the rotor packet. Then, theassembly of the rotor shaft 427 and the output shaft 426 with thecomponents mounted thereon is inserted from the output side into themotor gear unit.

The second ball bearing 242 and the output side inner gear 434 areplaced on the output side eccentric disk 429 from the output side. Thering 441 is placed on the centered circular disk 430 and the third ballbearing 425 is placed on the ring 441.

The carrier rollers 443 are placed on the carrier pins 442 of the outputring 440 and the output ring 440 is placed on the third ball bearing425, such that an inner shoulder of the output ring 440 rests on theouter ring of the third ball bearing 425. In this case, the carrier pins442 and the carrier rollers 443 are guided through the circular openings444 of the output side inner gear 433 and the drive side inner gear 434.The ball bearing 31 is mounted to a shoulder of the output pulley andthe transmission cover is screwed together with the outer gear to themotor housing 22, wherein the fastening openings of the transmissioncover and the outer gear overlap.

The rotor packet 26 and the inner rotor shaft 27′ are rotated byenergizing the stator 22. This rotation is transmitted to the eccentricdisks 428, 429. The eccentric disks 428, 429 in turn set the inner gears433, 434 disposed thereon in an eccentric circular movement, whereby thetoothings of the inner gears 433, 434 are moved past the toothing of thestationary outer gear 439. As a result, the internal toothing 437 of theouter gear 439 exert a reaction force on the inner gears 433, 434. Bythis reaction force of the outer gear 439, the inner gears 433, 434 areset in rotation about their own axis of rotation. This rotation istapped by the carrier pins 442 and transmitted to the output pulley,whereby the movement of the inner gears 433, 434 is converted into acentered circular motion. From there, the rotation is transmitteddirectly to a load, such as in a geared motor, or, as for example in thecase of an electric bicycle, initially transmitted to an output shaft 39via a freewheel 40.

FIGS. 82-88 show a tension shaft transmission and a motor gear unit witha tension shaft transmission according to the present description.

FIG. 82 shows a cross-section of a motor gear unit with the tensionshaft or flexspline transmission according to the present description.The elements not shown here correspond to those of FIG. 1.

An external toothing 5″ of a tensioning shaft 453 is arranged between acam disk 455′ with a flexible ball bearing 33 and an outer gear 457 withan internal toothing 6″, which is screwed to a motor housing 22. Thetensioning shaft is fastened via a fastening region with rivets to anoutput shaft 458, which is mounted on the diagonally opposite ballbearings 31 and 30.

FIG. 83 shows an exploded view of the tension shaft or flexsplinetransmission of FIG. 82, in which a stationary outer gear 8″ is providedwith the toothing referred to below as toothing for the HPD-Ftransmission. As a result, a particularly good engagement of the twotoothings can be achieved.

Thereafter, the internal toothing of the outer gear is substantially anouter equidistant to the gear trajectory that is defined by

x(t)=r1*cos(t)+r2*cos((n+1)*t)+r3*cos((n+3)*t) and

y(t)=r1*sin(t)−r2*sin((n+1)*t)+r3*sin((n+3)*t), wherein t is between 0and 2 pi/Z_outer or 360°/Z_outer, respectively, and where theequidistant, for example, has a distance of a pin radius to the geartrajectory.

The opposite external toothing of the tensioning shaft is derived fromthe geometry of a pin ring with cylindrical pins. Thus, a cross sectionof the tooth tips in a plane perpendicular to the axial directioncorresponds to a sector of a circle, preferably a semicircle. This canbe seen in FIG. 87.

The exploded view of FIG. 83 shows, viewed from left to right or fromthe output side to the drive side, respectively, the stationary outergear 8″ or the outer ring 8″, the cup-shaped tensioning shaft 453 withthe internal toothing 6″ and the fastening region 454, the flexible ballbearing 33 and the drive cylinder 455 with oval circumference andfastening flange 456. The drive cylinder of FIG. 83 is suitable, forexample, for a geared motor and deviates from the cam disk 455′ of FIG.82, which is arranged on an inner rotor shaft.

In the present specification, “oval” preferably denotes an oval havingtwo mutually perpendicular mirror symmetries or major axes, such as anellipse or a sinusoidal superimposed circular shape. However, it mayalso, for example, designate an oval with three axes of symmetry, inwhich the distance of the axes is maximum, so that three engagementregions are generated instead of two engagement regions with completetooth engagement.

The fastening region 454 of the tensioning shaft is suitable forfastening an output shaft. Furthermore, the stationary outer gear 8″comprises a fastening region 457 for attachment to a transmissionhousing and the drive cylinder 455 comprises a flange 456 for fasteninga drive axle. In an embodiment of the tension shaft transmission for anelectric bicycle, the drive cylinder 455 may also be formed as part ofan inner rotor shaft and need not have a fastening flange 456 in thiscase.

FIG. 84 shows a side view of the tension shaft transmission of FIG. 82from the drive side in the assembled state, in which, from inside tooutside, the fastening region of the tensioning shaft, the drivingcylinder, the flexible ball bearing, the external toothing of thetensioning shaft, the internal toothing of the outer gear and the outergear are shown.

FIG. 85 shows a cross-section along the cross-sectional line A-A of FIG.84, which runs along a semiminor axis of the drive cylinder 455. In thecross-sectional view of FIG. 85, it can be seen that the tensioningshaft 453 is biased inwards, so that it rests against the outer ring ofthe flexible ball bearing 33 in the region of the semiminor axis of thedrive cylinder 455.

FIG. 86 shows a cross-section along the cross-sectional line B-B of FIG.84, which extends along a semi major axis of the drive cylinder 455. Asshown in the cross-sectional views of FIGS. 85 and 86, the outer gear 8″and the tensioning shaft 453 each comprise smooth, cylindrical holes oropenings, whereas the drive cylinder 455 comprises threaded holes.

FIG. 87 shows an enlarged detail designated “C” in the previous FIG. 84,in which the shape of the eccentric toothing can be seen.

FIG. 88 shows the tension shaft transmission of FIG. 82 in the assembledstate.

FIGS. 89-93 show a two-stage reduction gear with a two-part one-piecepin ring 102′ and two outer gears 8′″, 8 ⁽⁴⁾, wherein the two-partone-piece pin ring 102′ is mounted on an eccentric disk.

The two-part one-piece pin ring 102 comprises a first external toothing5′″ on a first portion and a second external toothing 5 ⁽⁴⁾ on a portionarranged next to it. The first external toothing 5′″ of the pin ring102′ is arranged opposite a first internal toothing 6′″ of a rotatablymounted outer gear 8″ and the second external toothing 5 ⁽⁴⁾ of the pinring 102′ is arranged opposite a second internal toothing 6 ⁽⁴⁾ of astationary outer gear 8 ⁽⁴⁾.

A number of teeth of the first external toothing 5′″ is less than anumber of teeth of the internal toothing 6″ and a number of the teeth ofthe second external toothing 5 ⁽⁴⁾ is less than a number of the teeth ofthe internal toothing 6 ⁽⁴⁾. In an exemplary embodiment of ahigh-reduction gear, a number of the teeth of the first externaltoothing 5′″ is equal to 28 and a number of the teeth of the internaltoothing 6′″ of the movable outer gear 8′″ is equal to 29, and a numberof the teeth of the second external toothing is equal to 29 and a numberof the teeth of the internal toothing 6 ⁽⁴⁾ of the stationary outer gearis equal to 30.

Thus, in this embodiment, a reduction of a drive side first gear stageis greater than a reduction of an output side second gear stage and anumber of teeth of the respective outer gear is greater than a number ofteeth of the respective radially opposite toothing of the pin ring. As aresult, the inner external toothing moves in the reference frame of theouter internal toothing opposite to the drive. Thus, the internaltoothing of the movable outer gear moves in the direction of the driveand reduces the reduction of the first gear stage. In a general case,the number of teeth of opposing toothings may be different.

When using an eccentric disk, the difference in the number of teeth mustbe at least one. Furthermore, the ball bearing and a transmissionelement arranged thereon, such as a pin ring, need not be deformable. Inan alternative embodiment, which comprises an oval disk instead of aneccentric disk and a flexible ball bearing, which is arranged on theeccentric disk, the difference of the numbers of teeth of opposingtoothings is a multiple of two.

In embodiments without a transmission means disposed between an innerand an outer gear, in which a transmission means simultaneously acts asan inner gear with external toothing, as for example in the tensionshaft transmission of FIG. 83 or in the two-stage eccentric gear of FIG.89 the difference of the numbers of teeth refers to the toothing of thetransmission means and the toothing opposite the transmission means.

FIG. 89 shows a side view of the two-stage reduction gear from theoutput side.

FIG. 90 shows a cross-section along the cross-sectional line A-A in FIG.89 in the viewing direction designated in FIG. 89. Thus, the upper halfof FIG. 90 shows a true cross section, whereas the lower half of FIG. 90shows a side plan view.

FIG. 91 shows a side view of the two-stage reduction gear of FIG. 82,wherein hidden components are indicated by dashed lines to illustratethe engagement of the opposing toothings.

FIG. 92 is a partially cutaway perspective view of the two-stagereduction gear of FIG. 82 viewed from the output side or the side of themovable outer gear 8′″, respectively.

FIG. 93 is a partially cutaway perspective view of the two-stagereduction gear of FIG. 82 viewed from the drive side or the side of thestationary outer gear 8 ⁽⁴⁾, respectively. The stationary outer gear 8⁽⁴⁾ comprises a fastening region for attachment to a transmissionhousing, which is not shown in FIGS. 89 to 93. Similarly, the movableouter gear 8′″ comprises a fastening region for fixing an output shaft,which is not shown in FIGS. 89 to 93.

The two-part pin ring of the two-stage pin ring transmission of FIGS.89-93 differs from an externally toothed tension shaft according to theprior art inter alia by the form of the toothing. The form of thetoothing of the two toothings pin rings is essentially an innerequidistant to a gear trajectory which is determined by formula (1)given below for a gear trajectory of a harmonic pin ring transmissionwith an eccentric disk, where the equidistant preferably has a distanceof a pin radius to the gear trajectory (1).

The harmonic pin ring transmission is also referred to as HPD-Etransmission. Formula (1) describes a non-retrograde epicyclictrajectory with one epicycle.

Correspondingly, the toothing of the stationary outer gear 8 ⁽⁴⁾ and ofthe movable outer gear 8″ is an external equidistant to the geartrajectory indicated by formula (1), preferably at a distance of a pinradius. The number of teeth is determined by the parameter “n” of theformula (1). It may suffice for both the external toothing and theinternal toothing if only the tooth heads are determined by formula (1).

Preferably, the toothings of the two-stage transmission of FIGS. 89-93are configured such that a highly geared transmission is formed in whicha reduction of the first gear stage is further reduced by the secondgear stage in that the second angular velocity generated by the secondgear stage is in opposite direction to a first angular velocitygenerated by the first gear stage. In addition, the second angularvelocity is less than twice the first angular velocity, so that theresulting angular velocity at the transmission output is smaller inmagnitude than the angular velocity of the first gear stage.

The second angular velocity can also be greater than the first angularvelocity, whereby a reversal of the direction of rotation is generated,so that the output now takes place in the same direction as the drive.The reversal of the direction of rotation can be favorable, for example,for a generator mode in a hybrid vehicle or else to superimpose therotational movements of drive and output.

In operation, the eccentric disk is rotated by a motor, for example anelectric motor. This rotational movement is converted via the ballbearing to an eccentric movement of the two-part pin ring. As a result,the teeth of the two-part pin ring are guided past the externaltoothings of the two outer gears or drawn in the external toothings.

This results in a rotational movement of the pin ring relative to thestationary outer gear and a rotational movement of the movable outergear relative to the moving pin ring, which overlap to an outputrotational movement.

The two-stage reduction gear can also be built with an oval transmitter,a deformable ball bearing and a deformable two-part pin ring. In thiscase, another internal toothing is preferably used than shown in FIGS.89-93. For example, the toothing geometry of the tension shafttransmission shown in FIGS. 83-87 may be used.

In particular, the two-stage reduction gear can also be operated in theopposite direction and used for a speed-up transmission. For example, ahigh gear ratio may be useful for a streak camera with a rotating mirrorthat can have several thousand revolutions per second.

A two-stage transmission gear with output outer gear can also be used inthe cycloidal gear. In this case, the output pulley with the carrierpins is eliminated. Instead, at least one inner gear is configured as asplit inner gear with two different toothings, wherein an internaltoothing of the rotatable outer gear is arranged opposite to the secondof the two toothings.

The description of the following FIGS. 94-118 discloses inner gear andouter gear toothings which may be used in conjunction with thetransmissions of the present description, in particular with theharmonic pin ring transmissions.

In particular, the present description discloses a harmonic pin ringtransmission with eccentric disk, which is also referred to as “HPD-Etransmission”. The HPD-E transmission comprises a first gear with afirst toothing, a second gear with a second toothing and a pin ring withround engagement regions.

The pin ring may be formed by a flexible pin retaining ring withcylindrical pins inserted therein. It may also be made in one piece,wherein the one-piece pin ring comprises an annular portion, from whichpin-like extensions protrude in the axial direction to the side.

In particular, when a third gear is provided as a support gear and whenthe pins or pin-like extensions are non-rotatably connected to theannular portion or the flexible pin retaining ring, the pins or pin-likeextensions on the side of the support gear may be configured only on oneside as a round engagement region.

For example, in a configuration having an outer gear and an inner gear,a second outer gear may be provided as a support gear concentricallyaligned with the first outer gear, wherein a radius and a toothing ofthe second outer gear coincides with the first outer gear and thetoothing of the second outer gear is aligned with the toothing of thefirst outer gear such that the tooth heads lie one behind the other inthe axial direction.

According to a further embodiment, the pin ring is formed in one pieceand has an internal toothing and an external toothing with round toothheads.

The round engagement region of the pin ring comprises a segment ofcircle like cross section and may generally be formed by the pins of thepin ring, the pin-like extensions of the pin ring or by round-shapedtooth heads of the pin ring. The segment of circle like cross section ispreferably unchanged along an axial direction of the transmission. Inparticular, the segment of circle like cross section may take the formof a semicircle or a full circle.

Furthermore, the HPD-E transmission comprises a revolving transmitterfor drawing the engagement regions of the pin ring into the firsttoothing of the first gear and into the second toothing of the secondgear. In this case, the first gear, the transmitter and the second gearare arranged concentrically with each other and the transmitter isarranged radially inside the pin ring.

The pin ring is arranged between the first gear and the second gear.When the first gear and the second gear are in the same axial plane, asis the case with an arrangement with an inner gear and an outer gear, anarrangement between the first gear and the second gear should beunderstood to mean that the round engagement regions, the pins or thepin-like extensions are arranged in the radial direction at leastpartially between the first gear and the second gear.

When the first gear and the second gear are in different axial plane, asis the case with an arrangement with two outer gears in which one of theouter gears is driven and the other outer gear is stationarily connectedto the housing, an arrangement between the first gear and the secondgear should be understood to mean that the pin-retaining ring or aportion of the pin ring corresponding to the pin retaining ring isarranged in the axial direction between the first gear and the secondgear.

The transmitter comprises a transmitter disk or cam disk arrangedeccentrically to a transmission central axis, wherein the disk may beconfigured in particular as a circular disk. Here, an annular structure,such as a ring which is attached via struts on a shaft of thetransmitter, is also considered as a disk. In operation, the transmitterdeforms the pin ring so that the outer and inner gears rotate relativeto each other.

The first toothing of the first gear and the second toothing of thesecond gear are respectively formed in accordance with an epicyclicconstruction which will be explained in more detail below with respectto an inner gear or an outer gear according to the epicyclicconstruction.

According to the epicyclic construction, locations on the respectivetooth surface of the first toothing or the second toothing are eachdetermined by a radial distance from the transmission central axis as afunction of a cycle angle.

The radial distance is in turn determined by an equidistant to a geartrajectory, wherein locations on the gear trajectory are each determinedby the vector sum of a cycle vector and an epicycle vector. In thiscase, a tail of the cycle vector lies on the transmission central axisand a tail of the epicycle vector lies in the tip of the cycle vector.

Furthermore, an epicycle angle of the epicycle vector is n times thecycle angle, and a length of the cycle vector is larger than a length ofthe epicycle vector, wherein n is a number of round engagement regionsof the harmonic pin ring transmission which is at least three.

This profile shape can also be summarized by the following formula,where the plus sign refers to the inner gear toothing and the minus signto the outer gear toothing.

$\overset{\rightarrow}{P} = {\begin{pmatrix}x_{p} \\y_{p}\end{pmatrix} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;\left( {n*\alpha} \right)}} \\{{r_{1}*{\sin(\alpha)}} \pm {r_{2}*\sin\;\left( {n*\alpha} \right)}}\end{pmatrix}}$

Depending on the design of the pin ring, the round engagement regiongiven above refers to the pins, to the pin-like extensions or to theround tooth heads of the toothings of the pin rings.

According to an embodiment, the first gear is an inner gear with anexternal toothing and the second gear is an outer gear with an internaltoothing. Thus, the first toothing is the external toothing of the innergear and the second toothing, the internal toothing of the outer gear.In this embodiment, the inner gear is disposed radially inside of theouter gear.

In the external toothing of the inner gear, the epicycle angle ismeasured in the same direction as the cycle angle and the equidistant isan inner equidistant. In contrast, in the internal toothing of the outergear, the epicycle angle is measured in the opposite direction to thecycle angle, and the equidistant is an outer equidistant. Thisconfiguration of the toothings will be explained in more detail belowwith respect to an inner gear and an outer gear, respectively.

According to a further exemplary embodiment of an HPD-E transmission,the first gear and the second gear are each an outer gear with aninternal toothing. Thus, the first toothing is formed by an internaltoothing of a first outer gear and the second toothing is formed by aninternal toothing of a second outer gear.

In the internal toothings of the two outer gears, the epicycle angle ismeasured in the opposite direction to the cycle angle and theequidistant is an outer equidistant.

In particular, the respective equidistant of the first toothing or ofthe second toothing may be an equidistant at a distance of the sum of aradius of the round engagement regions and a correction value, thecorrection value being dependent on a back lash. For cylindrical pins,the radius of the round engagement region is equal to the pin radius ofthe pins. For approximately cylindrical pins, the radius corresponds toa radius of a round engagement region of the pins. The correction valueis greater than or equal to zero, in particular it can also be equal tozero. If the correction value is greater than zero, the correctionfactor corresponds to a fraction of the radius of the round engagementregion, for example 5% or 10%.

According to an embodiment, the harmonic pin ring transmission maycomprise a rolling bearing, which rests on the transmitter disk, whereinthe cycle radius is equal to half the diameter of the rolling bearing.According to another embodiment, in which the round engagement regionsof the pin ring rest directly on the transmitter disk, the cycle radiusis equal to half the diameter of the transmitter disk. In this case, thetransmitter disk may be rotatably mounted on an inner side.

In particular, the epicycle radius may be equal to half the eccentricoffset by which the transmitter disk is offset with respect to thetransmission central axis.

Different combinations of drive and output are possible for the HPD-Etransmission. These options are available for both the HPD-Etransmission and the HPD-F transmission. The drive shaft may beconfigured in particular as a rotor of an electric motor.

In particular, a drive shaft may be connected to the transmitter. Inthis case, an output shaft may be connected to either the first gear,the second gear or the pin ring. When the output shaft is connected tothe first or the second gear, then the respective other gear isgenerally fixed to the housing or connected to the transmission housing.

When the output torque is received from the pins, this can for examplebe done by a can-shaped component comprising openings for inserting thepins. The can-shaped or cylindrical component can in turn be mounted onthe transmission housing for stabilization. In this case, usually eitherthe outer gear or the inner gear moves along with the pin ring, whilethe other gear is fixed to the transmission housing. It is also possiblein this case to omit the wheel, which moves along only with the pinring.

Furthermore, the present description discloses an inner gear for the“HPD-E transmission” with a single eccentric comprising a pin ring withround engagement regions. The inner gear comprises an external toothing,wherein geometric locations on the tooth surface of the externaltoothing are each determined by a radial distance from a central axis ofthe inner gear as a function of a cycle angle α (=symbol alpha).

The radial distance is in turn determined by an inner equidistant to agear trajectory. Geometric locations on the gear trajectory are eachdetermined by the vector sum of a cycle vector and an epicycle vector.In this case, a tail of the cycle vector lies on a central axis of theouter gear, and a tail of the epicycle vector lies in the tip of thecycle vector. Furthermore, the cycle vector and the epicycle vector orthe epicycle vectors are located in a common plane which isperpendicular to the central axis.

The cycle angle and an epicycle angle of the epicycle vector aredetermined relative to a reference line which extends perpendicular tothe central axis of the inner gear and through the central axis of theinner gear. This also applies to the cycle angles and epicycle anglesmentioned below, which are determined relative to a reference line whichextends perpendicular to the central axis of the respective gear andthrough the central axis of the respective gear. In the installed stateof the gear or the inner or outer gear, this central axis coincides withthe transmission central axis.

An epicycle angle of the epicycle vector is n times as large as thecycle angle, wherein the epicycle angle is measured in the samedirection as the cycle angle and wherein n is a number of pins of theharmonic pin ring transmission greater than two. A number Z_inner ofteeth of the inner gear is at least two and preferably one less than thenumber of pins. Thus, based on the number of teeth of the inner gear,the epicycle angle is (Z_inner+1) times as large as the cycle angle,with Z_inner being at least two.

A length of the cycle vector is greater than a length of epicyclevector. In particular, the radii or the lengths of the sum vectors maybe selected such that the gear trajectory is non-retrograde, that is tosay has no self-intersections.

This profile shape of the toothing of the inner gear can also besummarized by the following formula

$\overset{\rightarrow}{P} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;\left( {\left( {n - 1} \right)*\alpha} \right)} + {r_{3}*{\cos\left( {\left( {n - 3} \right)*\alpha} \right)}}} \\{{r_{1}*{\sin(\alpha)}} + {r_{2}*\sin\;\left( {\left( {n - 1} \right)*\alpha} \right)} - {r_{3}*{\sin\left( {\left( {n - 3} \right)*\alpha} \right)}}}\end{pmatrix}$

Furthermore, the present description discloses an outer gear for aharmonic pin ring transmission with a single eccentric wherein the pinspreferably have a circular cross-section. The outer wheel comprises aninternal toothing, wherein geometric locations on the tooth surface ofthe internal toothing are each determined by a radial distance from acentral axis of the outer gear as a function of a cycle angle α.

The radial distance is in turn defined by an outer equidistant to a geartrajectory, the terms “inner equidistant” and “outer equidistant” beingunderstood in relation to the distance from the respective central axisof the gear.

The geometric locations on the gear trajectory are each determined bythe vector sum of a cycle vector and an epicycle vector, wherein a tailof the cycle vector lies on a central axis of the outer gear, and a tailof the epicycle vector lies in the tip of the cycle vector.

Furthermore, an epicycle angle of the epicycle vector is n times aslarge as the cycle angle, and the epicycle angle is measured in theopposite direction to the cycle angle, wherein n is a number of pins ofthe harmonic pin ring transmission greater than two. The number Z_outerof the teeth of the outer gear is preferably one greater than the numberof pins. Thus, based on the number Z_outer of the teeth of the outergear, the epicycel angle is (Z outside−1) times as large as the cycleangle, with Z_outer being at least four.

A length of the cycle vector is greater than a length of epicyclevector. In particular, the lengths of the vectors or the ratio of thevector lengths may be selected such that the gear trajectory has noself-intersections.

In the epicyclic construction, the length of the cycle vector determinesthe mean distance of the toothing from the central axis, i.e. the pitchcircle while the length of the epicycle vector or the epicycle vectorsdetermines the height of the teeth.

Preferably, the tooth geometry, i.e. the radial distance from thecentral axis of the gear, is independent of a position in the axialdirection on the central axis. If pins are provided, a cross section ofthe pins is preferably independent of a position on the longitudinalaxis of the pins. The cross section of the pins is preferably circular,but it may be other than circular. For example, the pins may comprise aslightly larger diameter in a circumferential direction of the pin ringthan perpendicular thereto. Accordingly, a cross section of the roundengagement regions is preferably independent of an axial position.

This profile shape of the toothing of the outer gear can also besummarized by the following formula:

$\overset{\rightarrow}{P} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;\left( {\left( {n + 1} \right)*\alpha} \right)} + {r_{3}*{\cos\left( {\left( {n + 3} \right)*\alpha} \right)}}} \\{{r_{1}*{\sin(\alpha)}} - {r_{2}*\sin\;\left( {\left( {n + 1} \right)*\alpha} \right)} + {r_{3}*{\sin\left( {\left( {n + 3} \right)*\alpha} \right)}}}\end{pmatrix}$

In another aspect, the present description discloses an inner gear for aharmonic pin ring transmission with an oval transmitter or a doubleeccentric, wherein oval also refers to oval-like designs, for examplewith an elliptical shape. This transmission is also referred to as“HPD-F transmission”.

The inner gear comprises an external toothing, wherein a tooth surfaceof the external toothing is determined by a radial distance from acentral axis of the outer gear as a function of a cycle angle α. Theradial distance from the central axis is in turn determined by an innerequidistant to a gear trajectory.

A geometric location on the gear trajectory is determined by the vectorsum of a cycle vector, a first epicycle vector and a second epicyclevector. A tail of the cycle vector lies on the central axis of the innergear, a tail of the first epicycle vector lies in the tip of the cyclevector, and a tail of the second epicycle vector lies in the tip of thefirst epicycle vector.

In a two epicycle design, the first epicycle vector is also referred toas a first order epicycle vector and the second epicycle vector is alsoreferred to as a second order epicycle vector. Accordingly, theassociated epicycles are also referred to as first or second orderepicycles, respectively. In a design with only one epicycle,accordingly, only a first order epicycle is present.

Furthermore, an epicycle angle of the first epicycle vector is (n−1)times as large as the cycle angle and an epicycle angle of the secondepicycle vector is (n−3) times as large as the cycle angle. Here, n is anumber of pins of the harmonic pin ring transmission, which is greaterthan two, so at least three. The first epicycle angle is measured in thesame direction as the cycle angle, and the second epicycle angle ismeasured in an opposite direction to the cycle angle. The inner gearcomprises at least two teeth and preferably comprises two teeth lessthan the number of pins.

Thus, based on the number of teeth Z_inner of the inner gear, the firstepicycle angle is (Z_inner+2−1)=(Z_inner+1) times as large as the cycleangle and the second epicycle angle is (Z_inner+2−3)=(Z_inner−1) timesas large as the cycle angle, where Z_inner is at least two.

Here, a length of the cycle vector is greater than the sum of thelengths of the first epicycle vector and the second epicycle vector, anda length of the first epicycle vector is greater than a length of thesecond epicycle vector. In particular, the ratios of the lengths orradii should be chosen such that the gear trajectory has noself-intersections.

Furthermore, the present description discloses an outer gear for aharmonic pin ring transmission with an oval eccentric or a doubleeccentric or for a HPD-F transmission. The outer gear comprises aninternal toothing, wherein geometric locations on the tooth surface ofthe internal toothing are each determined by a radial distance from acentral axis of the outer gear as a function of a cycle angle α.

The radial distance is in turn defined by an outer equidistant to a geartrajectory, wherein geometric locations on the gear trajectory are ineach case determined by the vector sum of a cycle vector, a firstepicycle vector and a second epicycle vector.

A tail of the cycle vector lies on the central axis, a tail of the firstepicycle vector lies in the tip of the cycle vector, and a tail of thesecond epicycle vector lies in the tip of the first epicycle vector.

Furthermore, an epicycle angle of the first epicycle vector is (n+1) istimes as large as the cycle angle, and an epicycle angle of the secondepicyclic vector is (n+3) times as large as the cycle angle. Here, n isa number of pins of the harmonic pin ring transmission, which is greaterthan three, that is at least four.

Preferably, the number Z_outer of the teeth of the outer gear is largerby two than the number of pins. Thus, based on the number Z_outer of theteeth of the outer gear, the first epicycle angle is(Z_outer−2+1)=(Z_outer−1) times as large as the cycle angle, and thesecond epicycle angle is (Z_outer−2+3)=Z_outer+1 times as large as thecycle angle, where Z_outer is at least 2+2+2=six in order to expressthis relationship with the minimum numbers.

The first epicycle angle is measured in an opposite direction to thecycle angle, and the second epicycle angle is measured in the samedirection as the cycle angle. Furthermore, a length of the cycle vectoris greater than the sum of the lengths of the first epicycle vector andthe second epicycle vector, and a length of the first epicycle vector isgreater than a length of the second epicycle vector. Preferably, theratios of the lengths or radii is chosen such that the gear trajectoryhas no self-intersections.

Further, the present description discloses a harmonic pin ringtransmission with an inner gear with the external toothing previouslydescribed with respect to the HPD-F transmission and an outer gear withthe internal toothing previously described with respect to the HPD-Ftransmission.

Furthermore, the transmission comprises a pin ring with round engagementregions. In particular, the round engagement regions can be formed bypins or pin-like extensions, which preferably comprise a circularcross-section. A revolving transmitter is provided for drawing roundengagement regions of the pin ring in the internal toothing of the outergear and in the external toothing of the inner gear. The inner gear, thetransmitter and the outer gear are arranged concentrically with eachother and the transmitter is arranged radially inside the pin ring.

The pin ring or are the round engagement regions or the pins or pin-likeextensions of the pin ring are arranged radially between the inner gearand the outer gear. The transmitter comprises an oval-shaped cam disk ora double eccentric. In operation, the cam disk or the double eccentricdeforms the pin ring so that the outer gear and the inner gear rotaterelative to each other.

In the configuration with a pair of an inner gear and an outer gear, theinner gear and the outer gear are arranged in a first axial plane. Aportion of the pin ring, which corresponds to a pin retaining ring andwhich is in contact with the transmitter, is arranged in a second axialplane. Expediently, a further outer gear is provided in a third axialplane for supporting the pin ring, wherein the second outer gear hassubstantially the same dimensions and the same tooth shape as the firstouter gear and wherein the toothing is aligned with that of the firstouter gear. The second axial plane lies between the first and the thirdaxial plane. These designs apply to both the HPD-E and the HPD-Ftransmission.

In another embodiment, the present description discloses a harmonic pinring transmission having a first outer gear according to the previouslywith respect to the HPD-F described internal toothing and a second outergear according to the previously with respect to the HPD-F describedinternal toothing.

In addition, the harmonic pin ring transmission comprises a pin ringwith round engagement regions and a revolving transmitter for drawingthe round engagement regions of the pin ring in the internal toothing ofthe first outer gear and in the internal toothing of the second outergear.

In this transmission, the transmitter, the first outer gear and thesecond outer gear are arranged concentrically with each other, and thetransmitter is arranged radially inside the pin ring. Other than in theabove-described pair of inner gear and outer gear, the first outer gearand second outer gear are in different axial planes, and the pin ring,or a middle portion of the pin ring corresponding to a pin retainingring, is arranged in the axial direction between the first outer gearand the second outer gear.

As with the HPD-E transmission, the HPD-F transmission also has variousoptions for connecting a drive shaft and an output shaft.

Among others, a drive shaft may be connected to the transmitter. In thiscase, an output shaft may be connected to the pin ring. Furthermore, inthe configuration with a pair of inner gear and outer gear, an outputshaft may be connected to the inner gear or to the outer gear.

As with the HPD-E transmission described above, there are also variouspossibilities in the previously described HPD-F transmission to connecta drive shaft and an output shaft, wherein the drive shaft may beconfigured in particular as a rotor of an electric motor.

In the configuration with two outer gears, in particular one of theouter gears may be driven off and the other fixed to the housing. Tothis end, an output shaft may be connected to the output outer wheel. Inthe two outer gears configuration, there is no need to have an innergear facing an outer gear in the radial direction. The respectivenondriven off gears are expediently connected to the housing or fixed tothe housing.

In the configuration with two outer gears, the second outer gear servesto receive the rotational motion from the pin ring. In this case, thenumber of teeth of the second outer gear may correspond to the number ofround engagement regions of the pin ring in order to ensure a betterengagement. In this case, the pin ring slightly twists because thenumber of teeth of the first outer gear is larger than the number ofpins or of the round engagement regions of the pin ring.

Both in the arrangement with two outer gears and in the arrangement witha pair of inner gear and outer gear and a further outer gear, it isexpedient to use a pin ring which comprises a central region, of whichthe pins, the pin-like extensions or the round engagement regionsprotrude on two opposite sides in the axial direction. In particular, ifthe pin ring is made of one piece, the round engagement regions mayhowever also extend continuously from one side to the other.

The respective equidistant of the epicyclic construction may be anequidistant at a distance of the sum of a radius of the round engagementregions and a correction value, the correction value being determined bya back lash. In this case, the round engagement regions may be formed inparticular by pins or pin-like extensions.

According to an embodiment, the transmitter of the previously describedharmonic pin ring transmission comprises an oval shaped cam disk and aflexible rolling bearing, wherein oval in particular also includesoval-like. Oval-like shapes are, for example, the sinusoidalsuperimposed circular shapes resulting from a Taylor expansion of theellipse equation.

The flexible rolling bearing rests on the oval shaped cam disk. Thecycle radius for the respective epicyclic construction of the toothshapes is equal to the sum of half a diameter of the flexible rollingbearing and a correction factor. The diameter of the flexible rollingbearing corresponds to a reference circle diameter of the pin ring orthe pin arrangement formed by the pin ring in the undeformed state.

According to a further embodiment, the harmonic pin ring transmissioncomprises a first circular disk arranged eccentrically to a transmissioncentral axis and a second circular disk arranged eccentrically to atransmission central axis. In this transmission, the cycle radius isequal to the sum of a mean radius of the envelope of the twoeccentrically arranged circular disks and a correction factor. Theenvelope is the curve which is formed by the inner circumference of atraction mechanism such as a pin ring when the traction mechanism isclamped onto the circular disks.

According to another embodiment, the first epicycle radius is less thanor equal to the sum of a half pin ring stroke and a second correctionfactor, the second correction factor being less than or equal to zero.The pin stroke is determined by the difference between the largest andthe smallest radius of the flexible rolling bearing in the deformedstate. In particular, the first epicycle radius may be greater than ¼ ofthe pin ring stroke, and in particular equal to ⅜ of the pin stroke.

The pin ring stroke is the stroke of the round engagement regions of thepin ring, i.e. the distance which the round engagement regions move inthe radial direction when the pin ring is deformed by the transmitter.

According to a further embodiment, the length of the second epicyclevector is about one third of the length of the first epicycle vector,where “about” may refer in particular to a range of plus or minus 10% or5%.

The tolerances of the transmission of the present description correspondto the usual technical tolerances for length dimensions and angulardimensions, such as DIN 7168 T1 or T2, or for toothing accuracy betweenpins and teeth, such as DIN 3961 or DIN 3976.

The tolerance values calculated according to a predetermined standardmay be used in particular to determine whether a given toothing matchesa toothing according to the present description within the tolerance.For this purpose, a normalized photographic view of a gear profile maybe used or a gear measuring machine, such as a flank testing devicewhich scans the gear surface mechanically or optically.

The tolerances can be related in particular to the tooth thickness or tothe center distance. The corresponding fit systems are also called “unitcenter distance” or “unit tooth thickness” fit system. For example, ameasured tooth profile may be considered coincident with a predeterminedtooth profile according to the present description if a distance from atooth flank does not exceed 5% or 1% of the predetermined tooththickness.

The distance can be measured, for example, perpendicular to the toothsurface or as a distance in the direction of a central axis of the gear.Compliance with the tolerance can also apply in the static sense, forexample, with a probability of 90% if a standard probabilitydistribution such as a Gaussian curve is used. This can also be takeninto account model-independent in that only a predetermined percentageof measuring points, for example 90%, needs to be within the tolerancelimit. Therefore, it is assumed that a sufficient number of measuringpoints is distributed sufficiently evenly over the tooth surface inorder to approximately detect the tooth profile.

The toothings are explained in more detail below with reference to thefollowing FIGS. 94-118.

FIGS. 94 and 95 show, by way of example, two types of transmissions ofharmonic pin ring transmissions, for which corresponding toothgeometries are disclosed in the present description.

FIG. 94 shows a harmonic pin ring transmission (HPRD-F) 510 with a camdisk and a deformable bearing resting thereon. The HPRD 510 comprises arotor 513, which is supported on a transmission housing via a ballbearing (not shown here). An outer ring or outer gear 508, which isarranged concentrically outside the rotor 513, comprises a first outergear toothing or outer toothing 506 formed as internal toothing on afirst side.

The outer ring 508 is attached to a cylindrical housing part 509. If theouter ring is driven off, this housing part in turn is rotatably mountedon the transmission housing. A second outer gear toothing or outertoothing 506′ formed as an internal toothing is formed on a second outerring 508′, which is inserted into the cylindrical housing part 509 on aside opposite the first side.

A first inner gear toothing or inner toothing 505 formed as an externaltoothing is formed on a periphery of an inner ring or inner gear 507 andarranged concentrically within the first outer toothing 506. Similarly,a second inner gear toothing or inner toothing 505′ formed as anexternal toothing is formed on a periphery of a second inner ring 507′and arranged concentrically within the second outer toothing 506′.

The inner toothing 505, 505′ and the outer toothing 506, 506′ arearranged concentrically to a transmission central axis, wherein theinner toothing 505, 505′ is rotatable about the transmission centralaxis. In other embodiments, the outer toothing 506, 506′ or the outerand the inner toothing may be rotatable about the transmission centralaxis or the inner toothing may be attached to the transmission housing,depending via which toothing the output or the drive takes place.

A flexible thin section ball bearing 502 is mounted on a speciallyshaped flange 504 of the rotor shaft 513. The flange 504 forms atransmitter and may be formed, for example, as an oval, oval-like orsinusoidal superimposed circular shape. Instead of a flange 504 formedon the rotor shaft 513, a correspondingly shaped disk or ring may beprovided on the motor shaft.

Instead of a cam disk and a flexible bearing applied thereto, thetransmitter in the transmission type of FIG. 94 may also have aso-called double eccentric, which is formed by two circular disksarranged eccentrically to the transmission central axis. The circulardisks may be rotatably mounted about their respective axis of symmetry,wherein the pin ring is mounted on the circular disks. Alternatively,the circular disks may be attached to the rotating transmitter and aflexible rolling bearing may be mounted on the circular disks, whereinthe pin ring rests on the roller bearing.

A flexible pin retaining ring 503 is disposed between the flexible thinsection ball bearing 502 and the outer toothing 506, 506′. On an innerside, the flexible pin retaining ring 503 comprises grooves forreceiving pins 501, which are arranged on the pin retaining ring 503 atregular intervals. The pins are cylindrically shaped and have a circularcross-section.

The pin retaining ring 503 is made flexible so that it can deform inaccordance with an angular position of the flange 504. Due to therigidity, the pin ring formed by the pin retaining ring 503 and the pins501 acts both as a pulling means, which pulls the output transmissionpart, and as a pressure means, which pushes the output transmissionpart.

FIG. 94 shows a three-row arrangement in which a first pair 507, 508 ofinner gear and outer gear is located in a first axial plane, a secondpair 507, 508′ of inner gear and outer gear is located in a second axialplane and a transmitter is located in a third axial plane locatedbetween the first axial plane and the second axial plane.

Likewise, a two and a half row arrangement with a pair of inner gear andouter gear and another outer gear is possible. This is advantageous, inparticular in the case of a driven inner gear, because then the outputof the second inner gear can generally not be led to the outside. Inthis case, in a three-row arrangement, the second inner gear only runsto support the pins.

For reasons of stability, it is advantageous if the pin ring issupported both inwardly and outwardly. However, both for an HPD-Ftransmission and for a HPD-E transmission, an arrangement with only twoouter gears or with only two inner gears, which are arrangedconcentrically to the transmission central axis, is also possible,wherein one gear is fixed to the transmission housing and the other gearis rotatable. In these arrangements, the pin ring is arranged axiallybetween the two inner gears or between the two outer gears and the pinsengage in the respective inner gears or outer gears.

FIG. 95 shows an exploded view of a harmonic pin ring transmission 510′with a single eccentric disk, also referred to for brevity as “HPD-E”transmission. Components similar to those of FIG. 94 have the samereference numerals or reference numerals with an apostrophe “′”. TheHPD-E transmission 510′ of FIG. 95 has a particularly simple design inwhich the inner gear comprises only 2 teeth and the outer gear comprises4 teeth. In general, the outer gear always comprises two teeth more thanthe inner gear in this type of transmission.

In the transmission arrangement of FIG. 95, a two-toothed inner gear 505is concentrically disposed within a first four-toothed outer gear 506′and concentric with a transmission central axis 515 in a first axialplane and a second outer gear 506 which is essentially identical to thefirst outer gear 506′, is arranged concentrically with the transmissioncentral axis 515 in a second axial plane. In a third axial plane lyingaxially between the first axial plane and the second axial plane, a disk514 arranged eccentrically on the transmission central axis, and a pinretaining ring 503 are arranged. The eccentrically arranged disk 514 isarranged on a shaft (not shown here), which may be, for example, therotor shaft of a motor.

The pin retaining ring 503 comprises three semicircular recesses 516,which at regular intervals on an inner circumference of the pinretaining ring 503. The three pins 501 are disposed in the semicircularrecesses 516 such that they protrude from the pin retaining ring 503 ontwo opposite sides in the axial direction and project into the first andsecond axial planes, respectively, so as to engage the outer geartoothings 506, 506′ and the inner gear toothing.

In a harmonic pin transmission with a single eccentric, the outer gearcomprises 2 teeth more than the inner gear and the number of pins is thearithmetic mean of the numbers of teeth. In the simplest case, thisresults in a transmission with 2 inner gear teeth, 3 pins and 4 outergear teeth, as shown in FIG. 95. In principle, the difference in thenumber of teeth can also be a multiple of two, but a smaller differencein the number of teeth results in higher reduction and better support ofthe torque.

A harmonic pin transmission with a single eccentric, in which a pin ringis pressed only at one blade location in the toothing of an outer gear,is hereinafter also referred to as “HPD-E transmission”. In extendedform, this term also refers to harmonic pin transmissions with multipleouter gears where only one such blade location is available per outergear. In particular, this is the case when a common eccentric isprovided for two or more outer gears, as in the embodiment of FIG. 2.

FIGS. 3 to 6 show epicyclic constructions for generating a toothgeometry according to the present description, wherein FIG. 96 refers tothe inner gear toothing of the HPD-E transmission, FIG. 97 refers to theouter gear tooting of the HPD-E transmission, FIG. 98 refers to theinner gear toothing of the HPD-F transmission and FIG. 99 refers to theouter gear tooting of the HPD-F transmission. These epicyclicconstructions are explained in more detail below.

In FIGS. 96 to 99, the cross section of a pin is symbolized by a circle520 and the position vector of the gear trajectory is designated by thereference numeral 521. The circular pin cross section has its center inthe position vector 521. When the cycle vector and the epicycle vectoror the cycle vector and the two epicycle vectors revolve at the angle αaccording to the angle indications of FIGS. 95 to 99, the pin crosssection 520 specifies two envelopes which define an inner and an outerequidistant to the gear trajectory and which determine the respectiveinner gear toothing and outer gear toothing of the transmission. In theangle indication, “n” denotes a pin number.

According to the present description, a gear trajectory is generated byan epicyclic construction, and from this gear trajectory in turn resultsthe tooth geometry of the HPD-E transmission as an envelope of thecircular pins or as equidistant, respectively.

The gear trajectory of an inner gear toothing of an HPD-E transmissionresults from the fact that a radius of a cycle rotates by 360 degreesabout the transmission axis, while a radius of an epicycle rotates inthe same direction by n*360 degrees relative to the reference system ofthe transmission axis, wherein n is the number of pins. On the otherhand, relative to the connecting line of the point of origin to thecenter of the epicycle, the epicycle radius rotates by (n−1)*360degrees, where n−1 is the number of teeth of the inner gear. The n−1revolutions of the epicycle radius result in n−1 maxima and n−1 minimaof the total radius, which correspond to the teeth and the recesses ortooth root surfaces in between.

Here, the radius r_1 of the cycle is equal to the radius of the bearingon which the pins rest plus the radius of the pins. This radius is alsocalled “first-order radius”. The radius r_2 of the epicycle is equal tohalf the pin radius. The tip of the epicycle describes the pintrajectory. An inner rolling curve results as the equidistant to the pintrajectory at a distance that results from the sum of the pin radius anda bearing clearance. This inner rolling curve is equal to the toothgeometry of the inner gear toothing.

The tooth shape of the outer gear toothing of the HPD-E transmissionresults from a similar construction whereby a radius of a cycle rotatesby 360 degrees about the transmission axis, while a radius of anepicycle rotates in the opposite direction by n*360 degrees. On theother hand, relative to the connecting line from the point of origin tothe center of the epicycle, the epicycle radius rotates by (n+1)*360degrees, where n+1 is the number of teeth of the inner gear. The n+1revolutions of the epicycle radius result in n+1 maxima and n+1 minimaof the total radius, which correspond to the teeth and the recesses ortooth root surfaces in between.

Expressed in formulas, the gear trajectory of the HPD-E transmission isdescribed by

$\overset{\rightarrow}{D_{1}D_{2}} = {{f\left( {r_{1};\alpha} \right)} = \begin{pmatrix}{r_{1}*{\cos(\alpha)}} \\{r_{1}*{\sin(\alpha)}}\end{pmatrix}}$$\overset{\rightarrow}{D_{2}P} = {{g\left( {r_{2};\beta} \right)} = {\begin{pmatrix}{r_{2}*{\cos(\beta)}} \\{{\pm r_{2}}*{\sin(\beta)}}\end{pmatrix} = \begin{pmatrix}{r_{2}*\cos\;\left( {n*\alpha} \right)} \\{{\pm r_{2}}*{\sin\left( {n*\alpha} \right)}}\end{pmatrix}}}$

wherein the plus sign refers to the inner gear toothing and the minussign to the outer gear toothing.

$\begin{matrix}{\overset{\rightarrow}{P} = {\begin{pmatrix}x_{p} \\y_{p}\end{pmatrix} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;\left( {n*\alpha} \right)}} \\{{r_{1}*{\sin(\alpha)}} \pm {r_{2}*\sin\;\left( {n*\alpha} \right)}}\end{pmatrix}}} & (1)\end{matrix}$

This is a parameter representation of the gear trajectory with theparameter a. Since the normal of the trajectory is perpendicular to thetangent, the equidistant at distance d can be obtained from a parametricrepresentation by the following formulas (2a) and (2b):

$\begin{matrix}{{x_{d}(\alpha)} = {{x(\alpha)} \pm \frac{d*{y^{\prime}(\alpha)}}{\sqrt{{x^{\prime}(\alpha)}^{2} + {y^{\prime}(\alpha)}^{2}}}}} & \left( {2\; a} \right) \\{{y_{d}(\alpha)} = {{y(\alpha)} \mp \frac{d*{x^{\prime}(\alpha)}}{\sqrt{{x^{\prime}(\alpha)}^{2} + {y^{\prime}(\alpha)}^{2}}}}} & \left( {2b} \right)\end{matrix}$

where the upper sign applies for the gear trajectory of the externaltoothing and the lower sign applies for the gear trajectory of theinternal toothing when the angle is run through counterclockwise. Here,the symbol x′ or y′ means the respective derivative with respect to theangle.

In the specific case of the HPD-E transmission, this results in:

${x_{d}(\alpha)} = {{{{r_{1}*\cos\;(\alpha)} \pm {r_{2}*\cos\;\left( {n*\alpha} \right)}} + \frac{d*\left( {{r_{1}*{\cos(\alpha)}} \pm {{nr}_{2}*{\cos\left( {n*\alpha} \right)}}} \right)}{\sqrt{r_{1}^{2} + {2{nr}_{1}{r_{2}\left( {{{\sin(\alpha)}{\sin\left( {n\;\alpha} \right)}} \pm {{\cos(\alpha)}\cos\;\left( {n\;\alpha} \right)}} \right)}} + {n^{2}r_{2}^{2}}}}} = {{{r_{1}*\cos\;(\alpha)} \pm {r_{2}*\cos\;\left( {n*\alpha} \right)}} + \frac{d*\left( {{r_{1}*{\cos(\alpha)}} \pm {{nr}_{2}*{\cos\left( {n*\alpha} \right)}}} \right)}{\sqrt{{r_{1}^{2} \mp {2{nr}_{1}r_{2}{\cos\left( {\left( {n \pm 1} \right)\alpha} \right)}}} + {n^{2}r_{2}^{2}}}}}}$${y_{d}(\alpha)} = {{{r_{1}*{\sin(\alpha)}} \mp {r_{2}*{\sin\left( {n*\alpha} \right)}}} - \frac{d*\left( {{{nr}_{2}*{\sin\left( {n*\alpha} \right)}} + {r_{1}*{\sin(\alpha)}}} \right)}{\sqrt{{r_{1}^{2} \mp {2{nr}_{1}r_{2}{\cos\left( {\left( {n \pm 1} \right)\alpha} \right)}}} + {n^{2}r_{2}^{2}}}}}$

The following boundary conditions apply to the HPD-E transmission:

n * d_(pin) < h(D₁; r₁) 0 < d_(pin) < r₁,

where h (D1, r1) is the perimeter length of the pin ring.

For example, for an HPD-E transmission, the following values of Table 1may be given concretely.

Variable Amount Unit Meaning/Purpose n 58 l Number of pins tkpin 103 mmReference circle diameter of the pin arrangement dpin 3 mm Diameter of apin Exver 0.75 mm Eccentric offset of the pin arrangement Bzinner 3 mmInner width of gear Bzouter 3 mm Outer width of gear Lpin 3 mm Length ofpin Gsi 0.0 mm Additional inner back lash between pin and gear Gsa 0.0mm Additional outer back lash between pin and gear

From the first four values n, tkpin, dpin and Exver, the followingderived values of Table 2 result:

Variable Amount Unit Formula Meaning/Purpose r1i 51.5 mm tkpin/2 Radiusof the cycle r1a 51.5 mm tkpin/2 Radius of the cycle r3 0.75 mm ExverRadius of the epicycle tpin 5.579 mm tkpin*π/n Pitch of the pins(tangential) Z_inner 148 l n − 1 Inner number of teeth Z_outer 152 l n +1 Outer number of teeth i1 37 l Z_inner/2 Transmission with fixed outergear i2 38 l Z_outer/2 Transmission with fixed inner gear

Similar to what has been done above for an HPD transmission witheccentric, for an HPD transmission with oval cam disk and deformablebearing, the tooth geometry can be obtained by an epicyclicconstruction, but by using first and second order epicycles. Such atransmission is also referred to as “HPD-F” transmission for the sake ofbrevity.

For the internal toothing of the HPD-F transmission, the gear trajectoryresults as a superposition of a cycle with radius r_1, a first orderepicycle with radius r_2 and a second order epicycle with radius r_3.Here, the first order epicycle rotates (n−1) times as fast in the samedirection as the cycle and the second order epicycle rotates (n−3) timesas fast and in opposite directions to the cycle.

$\overset{\rightarrow}{D_{1}D_{2}} = {{f\left( {r_{1};\alpha} \right)} = \begin{pmatrix}{r_{1}*{\cos(\alpha)}} \\{r_{1}*{\sin(\alpha)}}\end{pmatrix}}$$\overset{\rightarrow}{D_{2}D_{3}} = {{g\left( {r_{2};\beta} \right)} = {\begin{pmatrix}{r_{2}*{\cos(\beta)}} \\{r_{2}*{\sin(\beta)}}\end{pmatrix} = \begin{pmatrix}{r_{2}*{\cos\left( {\left( {n - 1} \right)*\alpha} \right)}} \\{r_{2}*{\sin\left( {\left( {n - 1} \right)*\alpha} \right)}}\end{pmatrix}}}$$\overset{\rightarrow}{D_{3}P} = {{h\left( {r_{3};\gamma} \right)} = {\begin{pmatrix}{r_{3}*{\cos(\gamma)}} \\{{- r_{3}}*{\sin(\gamma)}}\end{pmatrix} = \begin{pmatrix}{r_{3}*{\cos\left( {\left( {n - 3} \right)*\alpha} \right)}} \\{{- r_{3}}*{\sin\left( {\left( {n - 3} \right)*\alpha} \right)}}\end{pmatrix}}}$

As a superposition of these three movements, the gear trajectory of theinner toothing of the HPD-F transmission results in:

$\begin{matrix}{{\overset{\rightarrow}{P} = {\begin{pmatrix}x_{p} \\x_{p}\end{pmatrix} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;(\beta)} + {r_{3}*\cos\;(\gamma)}} \\{{r_{1}*{\sin(\alpha)}} + {r_{2}*\sin\;(\beta)} - {r_{3}*{\sin(\gamma)}}}\end{pmatrix}}}{{and}\mspace{14mu}{thus}}{\overset{\rightarrow}{P} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*{\cos\left( {\left( {n - 1} \right)*\alpha} \right)}} + {r_{3}*{\cos\left( {\left( {n - 3} \right)*\alpha} \right)}}} \\{{r_{1}*{\sin(\alpha)}} + {r_{2}*{\sin\left( {\left( {n - 1} \right)*\alpha} \right)}} - {r_{3}*{\sin\left( {\left( {n - 3} \right)*\alpha} \right)}}}\end{pmatrix}}} & (3)\end{matrix}$

For the simplest case with 2 internal teeth, 4 pins and 6 externalteeth, the gear trajectory of the inner toothing results in:

$\overset{\rightarrow}{P} = {\begin{pmatrix}x_{p} \\x_{p}\end{pmatrix} = {\begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;\left( {3*\alpha} \right)} + {r_{3}*\cos\;(\alpha)}} \\{{r_{1}*{\sin(\alpha)}} + {r_{2}*\sin\;\left( {3*\alpha} \right)} - {r_{3}*{\sin(\alpha)}}}\end{pmatrix} = \begin{pmatrix}{{\left( {r_{1} + r_{3}} \right)*\cos\;(\alpha)} + {r_{2}*{\cos\left( {3*\alpha} \right)}}} \\{{\left( {r_{1} - r_{3}} \right)*\sin\;(\alpha)} + {r_{2}*{\sin\left( {3*\alpha} \right)}}}\end{pmatrix}}}$

For the values of r_1, r_2 and r_3, the following values may be used inparticular: r_1 can correspond to half the diameter or half the diameterof the reference circle of the undeformed pin arrangement, r_2 cancorrespond to three eighths of the pin stroke and r_3 to one third ofthe radius r_2. The pin stroke is again given by the difference betweenthe largest and the smallest radius of the cam disk.

According to another embodiment, correction terms are inserted, whichare determined by parameters a and b. According to this correction, theradius r_1 is replaced by the effective radius r_1+(a+b)/2 and theradius r_2 is replaced by the effective radius r_2−(b−a)/2. Taking thiscorrection into account results in:

$\begin{matrix}{{{\overset{\rightarrow}{P}\left( {x,y,a,b} \right)} = \begin{pmatrix}{{r_{1,{eff}}*\cos\;(\alpha)} + {r_{2,{eff}}*{\cos\left( {\left( {n - 1} \right)*\alpha} \right)}} + {r_{3}*{\cos\left( {\left( {n - 3} \right)*\alpha} \right)}}} \\{{r_{1,{eff}}*{\sin(\alpha)}} + {r_{2,{eff}}*{\sin\left( {\left( {n - 1} \right)*\alpha} \right)}} - {r_{3}*{\sin\left( {\left( {n - 3} \right)*\alpha} \right)}}}\end{pmatrix}}{wherein}{r_{1,{eff}} = {r_{1} + {\left( {a + b} \right)/2}}}{and}{r_{2,{eff}} = {r_{2} - {\left( {b - a} \right)/2.}}}} & (4)\end{matrix}$

Like with the HPD-E transmission mentioned above, the tooth geometryresults as an equidistant to the gear trajectory at the distance of thepin radius, to which a positive correction factor may be added toaccount for a back lash.

Accordingly, external toothing results as a superposition of a cyclewith radius r_1, a first order epicycle with radius r_2 and a secondorder epicycle with radius r_3. Here, the first order epicycle rotates(n+1) times as fast and in opposite directions to the cycle and thesecond order epicycle rotates (n+3) times as fast in the same directionas the cycle.

Thus, the following formula results for the gear trajectory of theexternal toothing of the HPD-F transmission:

$\begin{matrix}{\overset{\rightarrow}{P} = {\begin{pmatrix}x_{p} \\y_{p}\end{pmatrix} = {\begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;(\beta)} + {r_{3}*\cos\;(\beta)}} \\{{r_{1}*{\sin(\alpha)}} - {r_{2}*\sin\;(\beta)} + {r_{3}*{\sin(\gamma)}}}\end{pmatrix} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {r_{2}*\cos\;\left( {\left( {n + 1} \right)*\alpha} \right)} + {r_{3}*\cos\;\left( {\left( {n + 3} \right)*\alpha} \right)}} \\{{r_{1}*{\sin(\alpha)}} - {r_{2}*\sin\;\left( {\left( {n + 1} \right)*\alpha} \right)} + {r_{3}*{\sin\left( {\left( {n + 3} \right)*\alpha} \right)}}}\end{pmatrix}}}} & (5)\end{matrix}$

The above-mentioned correction terms also apply accordingly to the geartrajectory of the external toothing so that the following formularesults:

$\begin{matrix}{{\overset{\rightarrow}{P}\left( {x,y,a,b} \right)} = \begin{pmatrix}{{r_{1,{eff}}*\cos\;(\alpha)} + {r_{2,{eff}}*{\cos\left( {\left( {n + 1} \right)*\alpha} \right)}} + {r_{3}*{\cos\left( {\left( {n + 3} \right)*\alpha} \right)}}} \\{{r_{1,{eff}}*{\sin(\alpha)}} - {r_{2,{eff}}*{\sin\left( {\left( {n + 1} \right)*\alpha} \right)}} + {r_{3}*{\sin\left( {\left( {n + 3} \right)*\alpha} \right)}}}\end{pmatrix}} & (6)\end{matrix}$

The following conditions apply for the ratio of the epicycle radius andthe pin radius as well as for the ratio of the epicyclic trajectory ofthe first order epicycle to the pin diameter:

0 < r_(pin) < 2 * r₂ n * d_(pin) < f(r₁; r₂; α; β)

Thus, r_2 is always greater than half the pin radius and f is alwaysgreater than the n-fold pin diameter, which is a lower limit for thecircumference of the pin ring.

For example, the following values of Table 3 may be given concretely foran HPD-F transmission.

Variable Amount Unit Meaning/Purpose n 150 1 Number of pins tkPnom 111.5mm Reference circle diameter of the pin arrangement nominal (undeformed)dpin 1.4 mm Diameter of a pin Stroke 1.29 mm Stroke of the pins (fromtransmission center) aedFL 12.5 mm equidistant flexible bearing (outercontour to cam disk) bFL 13 mm Width flexible bearing kwKS 0 mmCorrection value cam disk (for compensation of backlash in flexible ballbearing and deformation of the cam disk) Bzinner 2.9 mm Inner width ofgear Bzouter 3 mm Outer width of gear Lpin 3 mm Length of pin a 0.03 mmTooth correction a, adaption with respect to major axis and shapingunder load b 0.07 mm Tooth correction b, adaption with respect to minoraxis and shaping under load Gsi 0.0 mm Additional inner back lashbetween pin and gear Gsa 0.0 mm Additional outer back lash between pinand gear

From the first five values n, tkPnom, dpin, Hub and aedFl of the abovetable, the following derived values of Table 4 result for the HPD-Ftransmission:

Variable Amount Unit formula Meaning/Purpose daLnom 110.1 mm tkPnom −dPin Outer diameter flexible bearing nominal, undeformed) diLnom 85.1 mmdaLnom − 2 Inner diameter aedfl flexible bearing (nominal, undeformed)KSA 86.39 mm diLnom + Stroke Cam disk, major axis “A” KSa 83.81 mmdiLnom − Stroke Cam disk, minor axis “a” r1 55.75 mm tkPnom/2 Radius ofthe cycle r2 0.484 mm Hub*3/8 Radius of the first order epicycle r30.161 mm r2/3 Radius of the second order epicycle tpin 2.335 mmtkPnom*π/n Pitch of the pins (tangential) Zinner 148 1 n − 2 Innernumber of teeth Zouter 152 1 n + 2 Outer number of teeth i1 37 1Zinner/4 Transmission with fixed outer gear i2 38 1 Zouter/4Transmission with fixed inner gear

A trajectory formed by the pin centers of the HPD-F transmission isderived from the sinusoidal superimposed circular form formed by thefollowing two vectors

$\overset{\rightarrow}{D_{1}D_{2}} = {{f\left( {r_{1};\alpha} \right)} = \begin{pmatrix}{r_{1}*{\cos(\alpha)}} \\{r_{1}*{\sin(\alpha)}}\end{pmatrix}}$$\overset{\rightarrow}{D_{2}M_{Pin}} = {{h\left( {{r_{2} + r_{3}};\beta} \right)} = \begin{pmatrix}{\left( {r_{2} + r_{3}} \right)*{\cos\left( {3\alpha} \right)}} \\{\left( {r_{2} + r_{3}} \right)*{\sin\left( {3\alpha} \right)}}\end{pmatrix}}$

The trajectory of the pin center points is obtained by adding the twovectors:

$\begin{matrix}{\begin{pmatrix}x_{M\_{Pin}} \\y_{M\_{Pin}}\end{pmatrix} = \begin{pmatrix}{{r_{1}*{\cos(\alpha)}} + {\left( {r_{2} + r_{3}} \right)*{\cos\left( {3\alpha} \right)}}} \\{{r_{1}*{\sin(\alpha)}} + {\left( {r_{2} + r_{3}} \right)*{\sin\left( {3\alpha} \right)}}}\end{pmatrix}} & (7)\end{matrix}$

For the HPD-E transmission, the trajectory of the pin center points is acircle with the radius r1, which is offset by half the pin stroke or bythe eccentric offset in relation to the transmission central axis 515.The respective trajectory of the pin center points rotates at the speedof the input shaft around the transmission central axis 515 with adriven transmitter. This also results in the trajectory of theindividual pins, which rotate around this rotating trajectory with theangular velocity of the output transmission part.

FIGS. 7-10 show the generation of a gear trajectory and an equidistantthereto for an inner gear toothing of an HPD-E transmission with 2 innergear teeth and 4 outer gear teeth. Similarly, FIGS. 11-14 show thegeneration of a gear trajectory and an equidistant thereto for an outergear toothing of the HPD-E transmission.

The tooth surface of the inner gear toothing is generated by anequidistant 519 to a gear trajectory 518 at a distance of half a pinradius. In FIGS. 100-103, the cross section of a pin is symbolized by acircle 520. Here, the inner equidistant, which is closer to thetransmission central axis 515 is meant by equidistant.

A location 21 on the gear trajectory 518 is given, as also shown in FIG.96, as the sum or superimposition of the cycle vector with the epicyclevector.

The tooth surface of the external gear toothing is generated by anequidistant 519 to a gear trajectory 518 at a distance of half the pinradius. In FIGS. 104-107, the cross section of a pin is symbolized by acircle 520. Here, the outer equidistant, which is farther away from thetransmission central axis 515 is meant by equidistant.

FIG. 108 shows a tooth geometry for an inner gear and an outer gearobtained for the HPD-E transmission with three pins using the followingparameters of Table 5.

Variable Amount Meaning dbearing 50 mm Diameter of the bearing dpin 20mm Pin diameter Exver  5 mm Eccentric offset Stroke 10 mm Pin stroke (=Exver x 2) Tkpin 70 mm Pin pitch circle

The cycle radius is set at 0.5*(dbearing+dpin)=35 mm and the epicycleradius is set equal to the eccentric offset Exver 5 mm.

FIGS. 16 to 18 show various views of the HPD-E transmission 510′ with 3pins shown in exploded view in FIG. 95.

FIG. 109 shows a top view of the HPD-E transmission 510′ from the sideof the eccentric 514.

FIG. 107 shows a cross-sectional view through the HPD-E transmission510′ along the cross-sectional line A-A shown in FIG. 16.

FIG. 111 shows a top view of the HPD-E transmission 510′ from the sideof the inner gear 507.

FIG. 112 shows another tooth geometry that results for the HPE-Etransmission with 58 pins with the parameters given above. Here, theinner gear toothing results from the equidistant 529 to the geartrajectory 528 of the inner gear and the outer gear toothing from theequidistant 529′ to the gear trajectory 528′ of the outer gear.

FIG. 113 shows another tooth geometry that results for the HPD-Ftransmission with 150 pins with the parameters given above. Here, forreasons of clarity, the associated gear trajectories are not shown, butonly the equidistants 529, 529′, which define the tooth geometry.

For illustration, two pins 1 of the total of 58 pins are shown in FIG.112 and in FIG. 113 three pins 1 of the total of 150 pins are shown. Therespective center point of the pins 1 is in the vicinity ofintersections of the gear trajectory 528 of the inner gear and the geartrajectory 528′ of the outer gear, as can be seen in FIG. 112.

FIG. 114 shows a gear trajectory 528 of the inner gear of thetransmission of FIG. 113 and the associated equidistant 529.

FIG. 115 shows a zero contour 525 of a toothing which was manufacturedaccording to an epicyclic construction without corrections, a worn outtoothing 527 after 147 hours of operation and a corrected toothing 526in which the effect of shrinkage is already taken into account by aprovision. During the running in of the transmission, the torque wastransmitted only in one direction, so that the wear took placepredominantly on the right-hand tooth flank. For better illustration,FIG. 115 contains additional auxiliary lines, such as the root circle orthe symmetry axis of the zero contour 525.

This provision can be taken into account, for example, by the correctionfactors a and b given above for the HPD-F transmission. The correctionis particularly relevant for the toothing of the HPD-F transmission, asit results in more pointed tooth shapes, which change more due to therounding off during the operation than the rounder tooth form of theHPD-E transmission

The one-sided stress on the tooth flanks, as in the example of FIG. 22,is typically given in a vehicle drive. If there is a one-sided stress,it may be useful to apply an asymmetric correction. In the epicyclicconstruction, this can be considered by an angle-dependent radius of anepicycle or the epicycles, wherein the epicycle radius is chosen suchthat for the inner gear toothing, a smaller radius is obtained and forthe outer gear toothing a larger radius is obtained when a startingangle is reached, wherein the starting angle lies before the toothcenter in the selected direction of rotation.

This radius is then continuously approximated to the radius of the zerocontour 525 until an end angle is reached, wherein the end angle can begiven in particular by the angular position of the tooth base. Forexample, the difference between the epicycle radius of the zero curveand the corrected curve may decrease to zero according to a linearfunction, a polynomial, a Gaussian curve, an exponential function, or anangular function.

In a two and a half row transmission arrangement with two outer gearsand only one inner gear, as shown for example in FIG. 95, a larger forceis transmitted to the inner gear than to the outer gear. As a result,more wear occurs on the inner gear. For this reason, it may be useful toprovide the correction only for the inner gear toothing. Furthermore, itmay be useful to manufacture the inner gear of a harder material thanthe two outer gears. For example, the inner gear may be made of steeland the outer gear or the outer gears of plastic. Suitable materialsinclude, for example, chromium-molybdenum steel and polyamide. FIG. 116shows a detail of an HPD-F transmission with a toothing according to theepicyclic constructions shown in FIGS. 98 and 99 and an integrallyformed pin ring 530. Unlike the embodiment of FIG. 113, the tractionmechanism is not formed by cylindrical pins 501 inserted in an elasticpin retaining ring 503 but by an integrally formed pin ring 530. Theintegrally formed pin ring 530 comprises an internal toothing andexternal toothing in which the tooth tips are round in a forcetransmitting region.

Instead of the pins 501, the integrally formed pin ring 530 comprises aninternal toothing and an external toothing, wherein the internaltoothing and the external toothing each comprise segment of circle likeouter regions or teeth 531 and segment of circle like inner regions ortooth bases 532, which are interconnected by transition regions 533.

A tooth base 532 of the outer toothing of the pin ring 530 is in eachcase opposite a tooth 531 of the inner toothing and a tooth 531 of theinner toothing is in each case opposite a tooth base 532 of the outertoothing. The number of teeth 531 of the outer and the inner toothingcorresponds to the number of pins in the embodiment of FIG. 113. Thus,the toothings of the pin ring 530 each comprise two teeth more than theinner gear and two teeth less than the outer gear.

The integral pin ring 530 is formed from an elastic material such asrubber, plastic or metal or a composite elastic material. In particular,it may be formed by milling from an annular element. The pin ring 530may also be used for an HPD-E transmission, wherein the number of teethis one greater than the number of teeth of the inner gear and onesmaller than the number of teeth of the outer gear.

FIG. 117 shows a tolerance range 535 defined by an inner envelope curve533 and an outer envelope curve 534 of a predetermined tooth profile536. According to a first measurement method, a tooth profile 537 to bemeasured is considered to coincide with the predetermined tooth profile536 if it lies within the envelopes 536, 537, which have the distance A(delta) from the predetermined tooth profile 536.

The distance A may, for example, be a fraction of a tooth width s, afraction of a tooth pitch p, a fraction of a distance r from a centralaxis of the gear or predetermined absolutely. In this case, the toothwidth s, for example, may be measured halfway between the tooth base andtooth tip or at the level of the inflection points of the profile curve536. Depending on the dimension of the toothing, the value delta can be,for example, 0.5 mm, 0.2 mm or 0.08 mm, or it can be, for example, 5%,1%, 0.5% or 0.2% of the tooth width s.

According to a further comparison method, a comparison value isdetermined by an average of the distances to the given profile atpredetermined measuring points, for example as an arithmetic mean valueor as a quadratic mean value. Depending on the dimension of thetoothing, for example, a comparison value of <=0.5 mm, <=0.2 mm or<=0.08 mm can be regarded as a good match with the predetermined profile536. The distance may be measured in particular perpendicular to thepredetermined profile curve 536. Further comparison methods can be takenfrom the industrial standards given above.

In particular, the predetermined tooth profile 536 may be a toothprofile given according to the present description, which is predefinedby the equations (1), (3) or (5) given above, for example. Equations(1), (3) and (5) describe an array of curves of profile curves fromwhich the profile curve is to be selected in which a comparison valuebetween the toothing to be measured and the selected profile curve isminimized under the given boundary conditions. For example, boundaryconditions, which are determined from the toothing to be measured, canalready specify the sum of the radii and the number n in the argument ofthe trigonometric functions.

In this approach, further deviations of the measured toothings comparedto the predetermined toothing, such as flank line deviations, pitchdeviations, and runout remain unconsidered. This can be achieved, interalia, by comparing only individual teeth in a given side view or byincluding such deviations in the measured deviation.

FIG. 118 shows a tolerance range 535′ of a tooth profile defined byprofile displacements. An outer curve 536′ is given by the profile 536displaced outward by the distance delta in the radial direction, and aninner curve 535′ is given by the profile 536 displaced inwards in theradial direction by the distance delta.

As with the example of FIG. 117, a tooth profile to be measured may beconsidered consistent with the given profile if it is located at allmeasurement points or statistically, for example for 95% of allmeasurement points, within the tolerance range defined by the outercurve 536′ and the inner curve 535′.

According to a bearing concept of the present description, which relatesin particular to the bearing of the rotor shaft, an output element andan output shaft connected thereto, three bearings are provided, on whichthe aforementioned components are mounted in the housing. In particular,the output element and the output shaft are configured such that theyare supported by two diagonally opposite bearings inwardly relative tothe rotor shaft and outwardly relative to the housing.

In addition, if there is a pedal shaft supported by two further bearingsin the rotor shaft, the three-bearing arrangement expands to afive-bearing arrangement. In this case, the hollow output shaftconnected to the output element forms an outer hollow output shaft,which is connected via an outer freewheel with a further hollow outputshaft. Furthermore, the pedal shaft is connected to the hollow outputshaft via an inner freewheel, wherein a speed-up transmission, such asby the planetary gears shown in FIGS. 67 to 76, may be switched betweenthe pedal shaft and the output shaft.

Advantageously, the inner gear and the output shaft may be made of onepiece by two interconnected hollow shafts, wherein the hollow outputshaft comprises a smaller diameter than the inner gear. An inner bearingmay then be disposed on an inner shoulder of the hollow output shaft anda diagonally opposite outer bearing may be disposed on an outer shoulderof the hollow output shaft, wherein the outer shoulder is located in theconnection region between the inner gear and the hollow output shaft.

In the present description, ball bearing may also be used generally fora rolling bearing. Roller bearings may be used in particular forbearings that serve for support. For the bearing of the reduction gear,which is arranged on the cam disk or eccentric disk, a ball bearing withballs is preferably used, but a roller bearing can also be used.

FIG. 119 shows a schematic drawing of a first drive.

In this drive, a drive torque is transmitted from a motor to a reductiongear and from there to an output shaft, wherein the reduction gear is atransmission of the types described in the present specification, asindicated for example in the schematic drawings 121-124.

In particular, the motor may be an electric motor such as a three-phaseexternal rotor motor. All of the reduction gears mentioned in thepresent description can also be installed in a motor gear unit without acrankshaft, as shown in FIG. 119, for example in a geared motor. In thiscase, the components arranged on the pedal shaft are eliminated. Such adesign is suitable for example for a robot arm.

FIG. 120 shows a schematic drawing of another drive.

In this drive, a driving torque is transmitted from a motor to areduction gear and from there to an outer freewheel of a doublefreewheel and to an output shaft.

Another drive torque is transmitted from a pedal shaft to an innerfreewheel of a double freewheel and to the output shaft. Instead of apedal shaft, also another type of crankshaft or generally a drive shaftmay be provided. In general, the inner drive shaft may also be driven byanother motor or by another mechanical drive such as wind or waterpower.

FIG. 121 shows a schematic drawing of a flexspline or tensioning shaftdrive according to the present description. In the flexspline drive, amotor torque is transferred to a transmitter and from there to aflexspline. The flexspline is rotatably mounted on the transmitter, forexample by a deformable ball bearing.

The flexspline is supported on a stationary outer gear, which is fixedto a housing. This support is indicated by a double arrow. The outergear and the housing thus absorb the reaction force of the output.

FIG. 122 shows a schematic drawing of a harmonic pin ring driveaccording to the present description. A motor power is transmitted to atransmitter and from there to a pin ring. The pin ring is rotatablymounted on the transmitter, for example by a deformable ball bearing.

The output torque is transmitted from the pin ring to a rotatablymounted inner gear and from there to an output shaft. The pin ring issupported on a stationary outer gear, which is fixed to a housing, whichis also indicated by a double arrow. The outer gear and the housing thusabsorb the reaction force of the output.

FIG. 123 shows a schematic drawing of an eccentric transmissionaccording to the present description. A motor torque is transmitted toone or more eccentrics and from there to one or more inner gears. Theinner gears are each rotatably supported by a bearing on the eccentrics.By means of an arrangement of bearings and journals, the torque istransmitted to an output plate.

The arrangement of bearings and journals represents a centering deviceor a centering transmission with a 1:1 ratio. In particular, twoeccentrics mutually offset by 180° may be provided.

FIGS. 124 to 126 show a further motor transmission unit with atensioning shaft transmission. Unlike the embodiment of FIG. 82, thetensioning shaft 453′ here is formed on an outer output shaft 458,rather than a tensioning shaft 453 being attached to an outer outputshaft 458 with rivets. The outer output shaft 458 with the tensioningshaft 435 is supported on diagonally opposite ball bearings 30 and 31.This corresponds to a three-bearing arrangement with the ball bearings30, 31 and 29 or a five-bearing arrangement with the ball bearings 30,31, 29, 45 and 46. Like the outer output shaft 458, the output shaft 39is also mounted in a space-saving manner on only two ball bearings 46,41, which are offset from each other diagonally.

Components which have already been described previously, for example inconnection with FIG. 1, will not be described again here and, forreasons of clarity, are generally not provided separately with referencenumerals.

The features of the embodiments of FIGS. 94-123 are also disclosed inthe features of the following listing, which may be combined with otherfeatures of the description. In particular, the aforementioned toothinggeometries can be combined with all transmissions of the presentdescription, wherein the toothing based on the epicycle gear trajectorywith two epicycles is preferably used in transmissions with an ovaltransmitter, and wherein the toothing based on the gear trajectory withone epicycle is preferably used in eccentric transmissions.

The dimensioning of opposing toothings and, if present, an intermediatetransmission means may be chosen according to the present description inparticular such that there is a complete tooth engagement, both in a pinring transmission as well as in a tension shaft transmission or in acycloidal gear, wherein the pin ring transmission and the tension shafttransmission can be built as a design with an eccentric transmitter oras a design with an oval transmitter.

-   1. Harmonic pin ring transmission, comprising    -   a first gear with a first toothing and    -   a second gear with a second toothing,    -   a pin ring with round engagement regions,    -   a revolving transmitter for drawing the engagement regions of        the pin ring in the first toothing of the first gear and in the        second toothing of the second gear,    -   wherein the first gear, the transmitter and the second gear are        arranged concentrically with each other and the transmitter is        arranged radially inside the pin ring and wherein the pin ring        is disposed between the first gear and the second gear, wherein        the transmitter comprises a transmitter disk disposed        eccentrically to a transmission central axis, wherein the first        toothing of the first gear and the second toothing of the second        gear are shaped according to an epicyclic construction,    -   where locations on the respective tooth surface of the first        toothing and the second toothing are each determined by a radial        distance from the transmission central axis as a function of a        cycle angle,    -   wherein the radial distance is in turn determined by an        equidistant to a gear trajectory, wherein locations on the gear        trajectory are respectively determined by the vector sum of a        cycle vector and an epicycle vector, wherein a tail of the cycle        vector lies on the transmission central axis and a tail of the        epicyclic vector lies in the tip of the cycle vector, wherein an        epicycle angle of the epicycle vector is n times as large as        that cycle angle and a length of the cycle angle is greater than        a length of the epicycle angle, wherein n is a number of the        round engagement portions of the harmonic pin ring transmission        which is at least three.-   2. Harmonic pin ring transmission according to item 1, wherein the    first gear is an inner gear with an external toothing and the second    gear is an outer gear with an internal toothing, wherein for the    external toothing of the inner gear, the epicycle angle is measured    in the same direction as the cycle angle and the equidistant is an    inner equidistant, and wherein for the internal toothing of the    outer gear, the epicycle angle is measured in the opposite direction    to the cycle angle and the equidistant is an outer equidistant.-   3. Harmonic pin ring transmission according to item 1, wherein the    first gear and the second gear are each an outer gear with an    internal toothing, wherein for the internal toothing of the two    outer gears, the epicycle angle is measured in the opposite    direction to the cycle angle and the equidistant is an outer    equidistant.-   4. Harmonic pin ring transmission according to item 3, wherein the    respective equidistant is an equidistant at a distance of the sum of    a radius of the round engagement regions and a correction value,    wherein the correction value depends on a back lash.-   5. Harmonic pin ring transmission according to item 3 or 4, wherein    the harmonic pin ring transmission comprises a rolling bearing which    rests on the transmitter disk, wherein the cycle radius is equal to    half the diameter of the rolling bearing.-   6. Harmonic pin ring transmission according to item 3 or 4, wherein    the cycle radius is equal to half the diameter of the transmitter    disk.-   7. Harmonic pin ring transmission according to any of items 3 to 6,    wherein the epicycle radius is equal to a half an eccentric offset    by which the transmitter disk is offset from the transmission    central axis.-   8. Harmonic pin ring transmission according to any of items 3 to 7,    wherein a drive shaft is connected to the transmitter.-   9. Harmonic pin ring transmission according to item 8, wherein an    output shaft is connected to the first gear.-   10. Harmonic pin ring transmission according to item 8, wherein an    output shaft is connected to the second gear.-   11. Harmonic pin ring transmission according to item 8, wherein an    output shaft is connected to the pin ring.-   12. Inner gear for a harmonic pin ring transmission with an external    toothing, wherein the tooth surface of the external toothing is    determined by a radial distance from a central axis of the inner    gear as a function of a cycle angle,    -   wherein the radial distance from the central axis is in turn        determined by an inner equidistant to a gear trajectory,    -   wherein a location on the gear trajectory is determined by the        vector sum of a cycle vector, a first epicycle vector and a        second epicycle vector, wherein a tail of the cycle vector lies        on the central axis, a tail of the first epicycle vector lies in        the tip of the cycle vector, and a tail of the second epicycle        vector lies in the tip of the first epicycle vector,    -   and wherein an epicycle angle of the first epicycle vector is        n−1 times as large as the cycle angle and an epicycle angle of        the second epicycle vector is n−3 times as large as the cycle        angle, wherein n is a number of pins of the harmonic pin ring        transmission which is at least four, wherein the first epicycle        angle is measured in the same direction as the cycle angle and        the second epicycle angle is measured in an opposite direction        to the cycle angle, and wherein a length of the cycle vector is        greater than the sum of the lengths of the first epicycle vector        and the second epicycle vector, and a length of the first        epicycle vector is greater than a length of the second epicycle        vector.-   13. Outer gear for a harmonic pin ring transmission with an internal    toothing, wherein locations on the tooth surface of the internal    toothing are each determined by a radial distance from a central    axis of the outer gear as a function of a cycle angle,    -   wherein the radial distance is in turn defined by an outer        equidistant to a gear trajectory,    -   wherein locations on the gear trajectory are each determined by        the vector sum of a cycle vector, a first epicycle vector and a        second epicycle vector, wherein a tail of the cycle vector lies        on the central axis, a tail of the first epicycle vector lies in        the tip of the cycle vector, and a tail of the second epicycle        vector lies in the tip of the first epicycle vector,    -   and wherein an epicycle angle of the first epicycle vector is        n+1 times as large as the cycle angle and an epicycle angle of        the second epicycle vector is n+3 times as large as the cycle        angle, wherein n is a number of pins of the harmonic pin ring        transmission which is at least four, wherein the first epicycle        angle is measured in an opposite direction to the cycle angle        and the second epicycle angle is measured in the same direction        as the cycle angle, and wherein a length of the cycle vector is        greater than the sum of the lengths of the first epicycle vector        and the second epicycle vector, and a length of the first        epicycle vector is greater than a length of the second epicycle        vector.-   14. Harmonic pin ring transmission, comprising    -   an inner gear according to item 12 and    -   an outer gear according to item 13,    -   a pin ring with round engagement regions,    -   a revolving transmitter for drawing the engagement regions of        the pin ring in the internal toothing of the outer gear and in        the external toothing of the inner gear, wherein the inner gear,        the transmitter and the outer gear are arranged concentrically        with each other, the transmitter is arranged radially inside the        pin ring and wherein the pin ring is disposed between the inner        gear and the outer gear.-   15. Harmonic pin ring transmission, comprising    -   a first outer gear according to item 13 and    -   a second outer gear according to item 13,    -   a pin ring with round engagement regions,    -   a revolving transmitter for drawing the engagement regions of        the pin ring in the internal toothing of the first outer gear        and in the internal toothing of the second outer gear,    -   wherein the transmitter, the first outer gear and the second        outer gear are arranged concentrically with each other, the        transmitter is arranged radially inside the pin ring and wherein        the pin ring is arranged in the axial direction between the        first outer gear and the second outer gear.-   16. Harmonic pin ring transmission according to item 14 or item 15,    wherein a drive shaft is connected to the transmitter.-   17. Harmonic pin ring transmission according to item 16, wherein an    output shaft is connected to the pin ring.-   18. Harmonic pin ring transmission according to item 14, wherein    -   a drive shaft is connected to the transmitter and an output        shaft is connected to the inner gear.-   19. Harmonic pin ring transmission according to item 14, wherein    -   a drive shaft is connected to the transmitter and an output        shaft is connected to the outer gear.-   20. Harmonic pin ring transmission according to item 15, wherein    -   a drive shaft is connected to the transmitter and an output        shaft is connected to one of the two outer gears.-   21. Harmonic pin ring transmission according to any of items 14 to    20, wherein the respective equidistant is an equidistant at a    distance of the sum of a radius of the round engagement regions and    a correction value, wherein the correction value is determined by a    back lash.-   22. Harmonic pin ring transmission according to any of items 14 to    21, wherein the transmitter comprises an oval shaped cam disk and a    flexible rolling bearing resting on the oval shaped cam disk,    wherein the cycle radius is equal to the sum of half the diameter of    the flexible rolling bearing and a correction value.-   23. Harmonic pin ring transmission according to any of items 14 to    21, wherein the transmitter comprises a first circular disk arranged    eccentrically to a transmission central axis and a second circular    disk arranged eccentrically to the transmission central axis,    wherein the cycle radius is equal to the sum of a mean radius of the    envelope of the two eccentrically arranged circular disks and a    correction value.-   24. Harmonic pin ring transmission according to any of items 14 to    23, wherein the first epicycle radius is less than or equal to the    sum of half a pin ring stroke and a second correction value, wherein    the second correction value is less than or equal to zero.-   25. Harmonic pin ring transmission according to any of items 14 to    24, wherein the length of the second epicycle vector is one third of    the length of the first epicycle vector.

The features of the embodiments of FIGS. 1-92 are also disclosed in thefeatures of the following listing, which may be combined with all otherfeatures of the description.

-   1. Harmonic pin ring transmission having an input shaft and an    output shaft, the transmission comprising:    -   a first outer gear, an inner gear concentrically disposed with        respect to the first outer gear in a first axial plane,    -   a second outer gear arranged in a second axial plane,    -   a traction mechanism extending between the first outer gear and        the inner gear,    -   a revolving transmitter which lifts the traction mechanism from        an outer circumference of the inner gear and presses it against        an inner circumference of the first outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein the cam disk is disposed in a third axial plane        located between the first axial plane and the second axial        plane, and    -   wherein the cam disk is formed in one piece with the hollow        drive shaft.-   2. Harmonic pin ring transmission according to item 1, wherein the    traction mechanism is formed as a pin ring, wherein pins protrude    from a middle portion on two opposite sides, wherein the middle    portion is disposed in the third axial plane, and wherein the    revolving transmitter lifts the pins from an outer circumference of    the inner gear and presses them against an inner circumference of    the first outer gear.-   3. Harmonic pin ring transmission according to item 1 or item 2,    wherein a circumference of the cam disk has an oval shape.-   4. Harmonic pin ring transmission according to item 1 or item 2,    wherein a circumference of the cam disk has a circular shape and is    arranged eccentrically to a transmission central axis.-   5. Harmonic pin ring transmission according to any of items 1 to 4,    wherein a rolling bearing is arranged between the cam disk and the    traction mechanism.-   6. Harmonic pin ring transmission according to any of items 1 to 5,    wherein the transmitter consists essentially of aluminum.-   7. Harmonic pin ring transmission according to any of items 1 to 6,    wherein the transmitter comprises a ring which is connected via    connecting struts with the hollow drive shaft.-   8. Harmonic pin ring transmission having an input shaft and an    output shaft, the transmission comprising:    -   a first outer gear, an inner gear concentrically disposed with        respect to the first outer gear in a first axial plane,    -   a second outer gear arranged in a second axial plane,    -   a traction mechanism extending between the first outer gear and        the inner gear,    -   a revolving transmitter which lifts the traction mechanism from        an outer circumference of the inner gear and presses it against        an inner circumference of the first outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein the cam disk is disposed in a third axial plane        located between the first axial plane and the second axial        plane, and wherein the first outer gear is formed by a first        outer ring and the second outer gear is formed by a second outer        ring, wherein the first ring and the second ring are inserted        into a support ring.-   9. Harmonic pin ring transmission according to item 8, wherein the    first outer ring and the second outer ring are each made of plastic.-   10. Harmonic pin ring transmission according to item 8 or item 9,    wherein the first outer ring and the second outer ring each comprise    radially outwardly projecting journals which are distributed over    the circumference of the respective outer ring, and wherein the    support ring comprises matching recesses into which the journals are    inserted.-   11. Harmonic pin ring transmission according to any of items 8 to    10, wherein the support ring is made of aluminum.-   12. Harmonic pin ring transmission according to any of items 8 to    11, wherein the support ring comprises two partial rings, which abut    in the axial direction.-   13. Harmonic pin ring transmission according to any of items 8 to    12, wherein the first outer ring, the second outer ring and the    support ring comprise screw holes matching each other.-   14. Harmonic pin ring transmission according to item 13, wherein    screws are passed through screw holes of a transmission cover and    through the matching screw holes of the first outer ring, the    support ring and the second outer ring and are screwed into a thread    of a transmission housing of the harmonic pin ring transmission.-   15. Harmonic pin ring transmission having an input shaft and an    output shaft, the transmission comprising:    -   a first outer gear, an inner gear concentrically disposed with        respect to the first outer gear in a first axial plane,    -   a second outer gear arranged in a second axial plane,    -   a traction mechanism extending between the first outer gear and        the inner gear,    -   a revolving transmitter which lifts the traction mechanism from        an outer circumference of the inner gear and presses it against        an inner circumference of the first outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein the cam disk is disposed in a third axial plane        located between the first axial plane and the second axial        plane,    -   and comprising a hollow output shaft which is mounted in the        inner gear via a motor freewheel, and a pedal shaft, which is        mounted in the hollow output shaft via a pedal shaft freewheel,        wherein the pedal shaft comprises a receiving region for the        motor freewheel on an outer circumference and comprises a        receiving region for the pedal shaft freewheel on an inner        circumference.-   16. Harmonic pin ring transmission according to item 15, wherein the    motor freewheel is configured as a clamp roller freewheel and the    pedal shaft freewheel is configured as a pawl freewheel.-   17. Harmonic pin ring transmission according to item 15 or item 16,    wherein the output shaft extends in the axial direction on the    output side of the hollow drive shaft, wherein a ball bearing is    arranged between the hollow output shaft and the pedal shaft, and    wherein the hollow output shaft comprises a fastening region for an    output element.-   18. Freewheel assembly having an outer clamp roller freewheel and an    inner pawl freewheel, comprising:    -   a hollow drive shaft,    -   a hollow output shaft,    -   a pedal shaft, wherein the pedal shaft, the hollow output shaft        and the hollow drive shaft are arranged concentrically to one        another, the hollow output shaft is arranged radially inside the        hollow drive shaft and the pedal shaft is arranged radially        inside the hollow output shaft,    -   wherein the hollow output shaft comprises a stair-shaped pawl        engagement region at an inner circumference and comprises a        stair-shaped clamp roller rolling region at an outer        circumference,    -   wherein the pedal shaft comprises a star-shaped receiving region        for pawls, wherein the star-shaped receiving region comprises        pawl seats and spring seats arranged adjacent to the pawl seats.-   19. Freewheel assembly according to item 18, wherein the    stair-shaped clamp body rolling region on the outer circumference of    the hollow output shaft and the stair-shaped pawl engagement region    on the inner circumference of the hollow output shaft are located in    essentially the same axial plane.-   20. Freewheel assembly according to item 18 or item 19, wherein the    hollow drive shaft comprises a disk-shaped region with an external    toothing, which is provided on an outer circumference of the    disk-shaped region.-   21. Freewheel assembly according to any of items 18 to 20, wherein    the hollow output shaft comprises an annular thickening at a first    end and comprises a fastening region for an output means at a second    end opposite the first end.-   22. Freewheel assembly according to any of items 18 to 21, wherein    the outer circumference of the hollow output shaft comprises a    bearing region for a ball bearing.-   23. Freewheel assembly according to any of items 18 to 22, wherein    the inner circumference of the hollow output shaft comprises a    bearing region for a ball bearing.-   24. Freewheel assembly according to any of items 18 to 23, wherein    the inner circumference of the hollow output shaft comprises an    internal thread at one end.-   25. Freewheel assembly according to any of the preceding items 18 to    24, further comprising:    -   pawls which are rotatably mounted in the pawl seats and spring        elements which are arranged in the spring seats and connected to        the pawls,    -   a freewheel cage with webs and clamp rollers arranged between        the webs, wherein the freewheel cage and the clamp rollers are        arranged radially between the clamp roller rolling region of the        hollow output shaft and an inner circumference of the hollow        drive shaft.-   26. Freewheel assembly according to any of items 18 to 25, wherein    the pawl seats are cylindrically shaped, are closed at one end by a    wall and are open at an opposite end.-   27. Freewheel assembly according to any of items 18 to 26, wherein    the stair-shaped clamp body rolling region and the freewheel cage    each comprise at least two receiving regions for spring elements and    wherein in each case a spring element is arranged between a    receiving region of the clamp body rolling region and a receiving    region of the freewheeling cage.-   28. Freewheel assembly according to any of items 18 to 27, wherein    the pedal shaft comprises a force sensor unit, wherein the force    sensor unit comprises a load cell and a pedal shaft ball bearing,    wherein the load cell is arranged on the pedal shaft ball bearing.-   29. Freewheel assembly according to item 28, wherein the load cell    comprises an inner annular portion which is attached via fastening    lugs to an outer annular portion, wherein the pedal shaft ball    bearing is inserted into the inner annular portion.-   30. Freewheel assembly according to item 29, wherein the inner    portion and the outer portion of the load cell are radially offset    to each other, wherein the fastening lugs are laterally bounded by    radial slots, and wherein at least two of the fastening lugs    comprise a strain gauge.-   31. Freewheel assembly according to item 29 or item 30, wherein an    axial thickness of the outer ring is reduced in the region of the    fastening lugs.-   32. Pedal shaft for a freewheel assembly, wherein the pedal shaft    comprises a first fastening region for a pedal crank at a first end    and a second fastening region for a pedal crank at a second end    opposite thereto,    -   and wherein the pedal shaft comprises a star-shaped receiving        region for pawls in the vicinity of the first end.-   33. Pedal shaft according to item 32, wherein the star-shaped    receiving region comprises steps, wherein the steps each comprise a    first side surface, a second side surface, a pawl support region    inclined in a predefined direction relative to the circumferential    direction and comprising a spring seat, an upper surface essentially    parallel to the circumference, and an end region with a pawl seat,    wherein the pawl seat is cylindrically shaped, axially open to one    side and closed to an axially opposite side.-   34. Harmonic pin ring transmission having an input shaft and an    output shaft, the transmission comprising:    -   a first outer gear, an inner gear concentrically disposed with        respect to the first outer gear in a first axial plane,    -   a second outer gear arranged in a second axial plane,    -   a traction mechanism extending between the first outer gear and        the inner gear,    -   a revolving transmitter which lifts the traction mechanism from        an outer circumference of the inner gear and presses it against        an inner circumference of the first outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein the cam disk is disposed in a third axial plane        located between the first axial plane and the second axial        plane,    -   and comprising a hollow output shaft which is mounted in the        inner gear via a motor freewheel, and a pedal shaft, which is        mounted in the hollow output shaft via a pedal shaft freewheel,        wherein the pedal shaft comprises a receiving region for the        motor freewheel on an outer circumference and comprises a        receiving region for the pedal shaft freewheel on an inner        circumference.-   35. Harmonic pin ring transmission according to item 34, wherein the    motor freewheel is configured as a clamp roller freewheel and the    pedal shaft freewheel is configured as a pawl freewheel.-   36. Harmonic pin ring transmission according to item 34 or item 35,    wherein the output shaft extends in the axial direction on the    output side of the hollow drive shaft, wherein a ball bearing is    arranged between the hollow output shaft and the pedal shaft, and    wherein the hollow output shaft comprises a fastening region for an    output element.-   37. Harmonic pin ring transmission having an input shaft and an    output shaft, the transmission comprising:    -   a first outer gear, an inner gear concentrically disposed with        respect to the first outer gear in a first axial plane,    -   a second outer gear arranged in a second axial plane,    -   a pin ring with pins extending between the first outer gear and        the inner gear,    -   a revolving transmitter which lifts the pins of the pin ring        from an outer circumference of the inner gear and presses them        against an inner circumference of the first outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein the cam disk is disposed in a third axial plane        located between the first axial plane and the second axial        plane,    -   and wherein the pin ring is formed in one piece and is made of        metal, wherein the pins are formed by projections which protrude        from two opposite sides of a central region of the pin ring,        wherein the central region comprises an inner bearing surface        for resting on a bearing.-   38. Harmonic pin ring transmission according to item 37, wherein the    projections are cylindrical on a first side of the two opposite    sides and wherein the projections are partially cylindrical on a    second of the two opposite sides, wherein a cylindrically shaped    region is located in the radial direction on the outside of the pin    ring.-   39. Harmonic pin ring transmission according to item 37, wherein the    projections comprise an inner rounded engagement region and an outer    rounded engagement region on a first side of the two opposite sides,    -   and wherein the projections comprise an outer rounded engagement        region on a second of the two opposite sides.-   40. Harmonic pin ring transmission according to any of items 37 to    39, wherein a bearing is arranged between the cam disk and the pin    ring, and wherein the pin ring comprises a shoulder on an inner side    for supporting the bearing.-   41. Pin ring for a harmonic pin ring transmission which is formed in    one piece and is made of metal, wherein the pin ring comprises pins    formed by projections protruding from two opposite sides of a    central region of the pin ring, wherein the central region comprises    an inner bearing surface for resting on a bearing outer bearing    surface.-   42. Pin ring according to item 41, wherein the projections are    cylindrical on a first side of the two opposite sides and wherein    the projections are partially cylindrical on a second of the two    opposite sides, wherein a cylindrically shaped region is located in    the radial direction on the outside of the pin ring.-   43. Pin ring according to item 41, wherein the projections comprise    an inner rounded engagement region and an outer rounded engagement    region on a first side of the two opposite sides,    -   and wherein the projections comprise an outer rounded engagement        region on a second of the two opposite sides.-   44. Pin ring according to item 41, wherein a web is located in each    case between the projections on the first side of the opposite two    sides, wherein an outer boundary line of the projections smoothly    merges into an outer boundary line of the web.-   45. Harmonic pin ring transmission according to any of items 41 to    44, wherein a bearing is arranged between the cam disk and the pin    ring, and wherein the pin ring comprises a shoulder on an inner side    for supporting the bearing.-   46. Harmonic pin ring transmission having an input shaft and an    output shaft, the transmission comprising:    -   a first outer gear, an inner gear concentrically disposed with        respect to the first outer gear in a first axial plane,    -   a second outer gear arranged in a second axial plane,    -   a traction mechanism extending between the first outer gear and        the inner gear,    -   a revolving transmitter which lifts the traction mechanism from        an outer circumference of the inner gear and presses it against        an inner circumference of the first outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein the cam disk is disposed in a third axial plane        located between the first axial plane and the second axial        plane,    -   wherein a pedal shaft is disposed radially inside the output        shaft, and wherein the pedal shaft is mounted via a drive side        pedal shaft ball bearing and a load cell in the motor housing.-   47. Harmonic pin ring transmission according to item 46, wherein the    load cell comprises an inner annular portion which is attached via    fastening lugs to an outer annular portion, wherein the pedal shaft    ball bearing is inserted into the inner annular portion and wherein    the outer annular portion is inserted into a cylindrical region of    the motor housing.-   48. Harmonic pin ring transmission according to any of items 46 to    47, wherein a wave spring is arranged between the load cell and the    drive side rotor ball bearing.-   49. Harmonic pin ring transmission having an input shaft and an    output shaft, the transmission comprising:    -   a first outer gear, an inner gear concentrically disposed with        respect to the first outer gear in a first axial plane,    -   a second outer gear arranged in a second axial plane,    -   a traction mechanism extending between the first outer gear and        the inner gear,    -   a revolving transmitter which lifts the traction mechanism from        an outer circumference of the inner gear and presses it against        an inner circumference of the first outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein the cam disk is disposed in a third axial plane        located between the first axial plane and the second axial        plane, wherein a pedal shaft is disposed radially inside the        output shaft and wherein further a planetary gear and a pedal        shaft freewheel are arranged in the flow of forces between the        pedal shaft and the output shaft.-   50. Harmonic pin ring transmission according to item 49, wherein a    planet carrier of the planetary gear is connected to the pedal    shaft, a ring gear of the planetary gear comprises a connection    region for connection to a transmission housing, a sun gear of the    planetary gear is mounted on the pedal shaft and wherein the pedal    shaft freewheel is disposed between a hollow shaft connected to the    sun gear and the output shaft.-   51. Harmonic pin ring transmission according to item 49, wherein the    pedal shaft freewheel is arranged between the crankshaft and a    planet carrier of the planetary gear, wherein a ring gear of the    planetary gear is rotatably mounted in the harmonic transmission and    wherein a sun gear of the planetary gear is adapted for attachment    to a stationary housing part.-   52. Tension shaft transmission, the tension shaft transmission    comprising:    -   an outer gear with an internal toothing, wherein the outer gear        comprises a fastening region for attachment to a transmission        housing,    -   a tensioning shaft having an external toothing, wherein the        tensioning shaft is concentrically disposed with respect to the        outer gear in an axial plane,    -   a revolving transmitter which presses the tensioning shaft        against the internal toothing of the outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk, wherein a ball bearing is arranged on a circumference of        the cam disk,    -   wherein a cross section of the tooth crests of the external        toothing of the tensioning shaft corresponds essentially to a        sector of a circle    -   and wherein, with respect to an axis of the outer gear, the        internal toothing of the outer gear is essentially defined by an        outer equidistant to the gear trajectory defined by the formulas        x(t)=r1*cost(t)+r2*cos((n+1)*t)+r3*cos((n+3)*t) and        y(t)=r1*sin(t)−r2*sin((n+1)*t)+r3*sin((n+3)*t), wherein n+1 is        the number of teeth of the internal toothing of the outer gear,        wherein the radii r1, r2 and r3 are greater than zero, and        wherein for the scale of the radii r2>r3 and r1>r2+r3 applies.-   53. Tension shaft transmission according to item 52, wherein the    tensioning shaft is formed in a cup shape, wherein at the bottom of    the cup shape, a fastening region is formed for fastening an output    shaft.-   54. Tension shaft transmission according to item 53, wherein a    central circular opening is formed at the bottom of the cup shape    and wherein the fastening region of the tensioning shaft comprises    fastening holes arranged around the central circular opening.-   55. Tension shaft transmission according to item 52, wherein the    tensioning shaft has the shape of a circular cylinder, and wherein    comprises a second outer gear which comprises a fastening region for    fastening an output shaft, wherein the internal toothing of the    outer gear is determined by the same construction as the internal    toothing of the first outer gear.-   56. Tension shaft transmission, the tension shaft transmission    comprising:    -   an outer gear with an internal toothing, wherein the outer gear        comprises a fastening region for attachment to a transmission        housing,    -   a tensioning shaft having an external toothing, wherein the        tensioning shaft is concentrically disposed with respect to the        outer gear in an axial plane,    -   wherein a cross section of the tooth crests of the external        toothing of the tensioning shaft corresponds essentially to a        sector of a circle,    -   a revolving transmitter which presses the tensioning shaft        against the internal toothing of the outer gear,    -   wherein the transmitter comprises a hollow drive shaft and a cam        disk,    -   wherein the tooth surface of the internal toothing of the outer        gear is determined by a radial distance from a central axis of        the inner gear as a function of a cycle angle, wherein the        radial distance from the central axis is in turn determined by        an inner equidistant to a gear trajectory, wherein locations on        the gear trajectory are each determined by the vector sum of a        cycle vector, a first epicycle vector and a second epicycle        vector, wherein a tail of the cycle vector lies on the central        axis, a tail of the first epicycle vector lies in the tip of the        cycle vector, and a tail of the second epicycle vector lies in        the tip of the first epicycle vector,    -   and wherein an epicycle angle of the first epicycle vector is        n+1 times as large as the cycle angle and an epicycle angle of        the second epicycle vector is n+3 times as large as the cycle        angle, wherein n is a number of pins of the harmonic pin ring        transmission which is at least four, and wherein a length of the        cycle vector is greater than the sum of the lengths of the first        epicycle vector and the second epicycle vector, and a length of        the first epicycle vector is greater than a length of the second        epicycle vector.-   57. Two-stage reduction gear, the two-stage reduction gear    comprising:    -   a stationary outer gear with a first internal toothing, wherein        the outer gear comprises a fastening region for attachment to a        transmission housing,    -   a rotatable outer gear with a second internal toothing, wherein        the outer gear comprises a fastening region for attachment to an        output shaft,    -   a two-part one-piece pin ring, wherein the two-part one-piece        pin ring comprises a first external toothing and a second        external toothing, wherein the first external toothing of the        two-part one-piece pin ring engages in the internal toothing of        the stationary outer gear and wherein the second external        toothing of the two-part one-piece pin ring engages in the        internal toothing of the rotatable outer gear,    -   a revolving transmitter which presses the two-part one-piece pin        ring against the internal toothing of the stationary outer gear        and against the internal toothing of the rotatable outer gear.-   58. Two-stage reduction gear according to item 57, wherein the    number of teeth of the internal toothing of the stationary outer    gear is greater than the number of teeth of the first external    toothing, and the number of teeth of the internal toothing of the    rotatable outer gear is greater than the number of teeth of the    second external toothing,    -   and wherein the number of teeth of the stationary outer gear is        greater than the number of teeth of the rotatable outer gear and        the number of teeth of the first external toothing is greater        than the number of teeth of the second external toothing.-   59. Two-stage reduction gear according to item 57 or item 58,    wherein the transmitter comprises a circular ring eccentric to the    axis of the stationary outer gear.-   60. Two-stage reduction gear according to any of items 57 to 59,    wherein a cross section of the tooth crests of the first external    toothing and a cross section of the tooth crests of the second    external toothing essentially correspond to a sector of a circle.-   61. Two-stage reduction gear according to any of items 57 to 60,    wherein a cross section of the tooth crests of the first external    toothing and a cross section of the tooth crests of the second    external toothing essentially correspond to an equidistant to the    gear trajectory defined by the formula x(t)=r1*cos(t)+r2*cos(nt) and    y(t)=r1*sin(t)+r2*sin(nt), wherein the following applies for the    radii r1, r2: r1>0, r2>0 and r1>r2.-   62. Two-stage reduction gear according to any of items 57 to 61,    wherein a cross section of the tooth crests of the first external    toothing and a cross section of the tooth crests of the second    external toothing essentially correspond to an equidistant to the    gear trajectory defined by the formula x(t)=r1*cos(t)+r2*cos(nt) and    y(t)=r1*sin(t)r2*sin(nt), wherein the following applies for the    radii r1, r2: r1>0, r2>0 and r1>r2.-   63. A load cell for determining a radial force acting on a    crankshaft, the load cell comprising:    -   a receiving sleeve for receiving a ring of a bearing,    -   a fastening ring for attaching the load cell in a transmission        housing,    -   axial support areas provided on the fastening ring for axially        supporting the ring of the bearing,    -   measuring regions for receiving radial forces of the receiving        sleeve and which connect the receiving sleeve with the fastening        ring, wherein strain sensors are attached to at least two of the        measuring regions.-   64. Freewheel assembly having an outer transmission freewheel and an    inner pedal shaft freewheel, comprising:    -   a hollow drive shaft,    -   a hollow output shaft,    -   a pedal shaft, wherein the pedal shaft, the hollow output shaft        and the hollow drive shaft are arranged concentrically with each        other, the hollow output shaft is disposed radially inside the        hollow drive shaft, and the pedal shaft is disposed radially        inside the hollow output shaft, wherein the pedal shaft        freewheel is arranged between the pedal shaft and the hollow        output shaft and wherein the transmission freewheel is arranged        opposite the pedal shaft freewheel on the hollow output shaft        and wherein the hollow output shaft comprises adapted areas on        an inner side and on an outer side in the region of the        respective freewheel.-   65. One-piece pin ring made of metal, wherein a pin retaining ring    and an arrangement of pins which protrude in axial direction from    the pin retaining ring on two opposite sides are made in one piece.-   66. One-piece pin ring according to item 65, wherein the pins are    connected to each other in the circumferential direction.-   67. One-piece pin ring according to item 65 or 66, wherein on a    first of the two opposite sides, the pins are formed as half pins,    which are adapted to engage an internal toothing, and wherein on a    second of the opposite side, the pins are formed as whole pins,    which are suitable for engaging in an internal toothing and for    engaging in an external toothing opposite the internal toothing.-   68. One-piece pin ring made of metal, wherein the one-piece pin    retaining ring comprises a smooth inner circumference on an inner    side and comprises rounded bulges on an outer side, which are made    in one piece with the pin retaining ring.-   69. One-piece pin retaining ring according to item 68, wherein at    least one head region of the rounded bulges comprises a semicircular    cross section.-   70. Support ring assembly for a reduction gear having a first outer    gear and a second outer gear comprising a support ring, a first    outer gear having a first internal toothing and a second outer gear    having a second internal toothing, wherein the first outer gear and    the second outer gear are inserted into the support ring on opposite    sides, and wherein the support ring comprises a fastening region for    attachment to a transmission housing.-   71. Support ring assembly according to item 70,    -   Wherein at least the first outer gear and the second outer gear        are made of plastic.-   72. Support ring assembly according to any of items 71 or 72,    wherein the first outer gear and the second outer gear are each    connected with the support wheel via a pin-groove connection.-   73. One-piece rotor-transmitter element for a reduction gear    comprising a hollow shaft comprising a fastening region on a first    side for fastening a rotor package, and comprising a cam disk on a    second side opposite to the first side, wherein an outer    circumference of the cam disk is configured as a receiving area for    a ball bearing.-   74. One-piece rotor-transmitter element according to item 73,    wherein the one-piece rotor-transmitter element is made of aluminum.-   75. One-piece rotor-transmitter element according to item 73 or 74,    wherein the hollow shaft is dimensioned such that a pedal shaft can    be passed through the hollow shaft.-   76. One-piece rotor-transmitter element according to any of items 73    to 75, wherein the cam disk comprises a circular circumference    arranged eccentrically relative to the central axis of the hollow    shaft.-   77. One-piece rotor-transmitter element according to any of items 73    to 75, wherein the cam disk comprises an oval circumference to the    central axis of the hollow shaft.-   78. Drive shaft with a planetary gear arranged on the drive shaft,    wherein a planet carrier of the planetary gear is fixedly connected    to the drive shaft, a ring gear of the planetary gear comprises a    fastening region for attachment to a transmission housing and a    receiving region for a torque sensor, and a sun gear of the    planetary gear is configured as a ring gear, which is arranged    concentrically to the drive shaft, and wherein the sun gear is    connected to a hollow output shaft of the planetary gear, which is    rotatably mounted on the drive shaft.-   79. Drive shaft with a planetary gear arranged on the drive shaft,    wherein a planet carrier of the planetary gear is mounted on the    drive shaft via a freewheel, a sun gear of the planetary gear    comprises a fastening region for attachment to a transmission    housing and a receiving region for a torque sensor, and a ring gear    of the planetary gear comprises a receiving region for a ball    bearing for supporting on a transmission housing.-   80. Drive shaft with a planetary gear arranged on the drive shaft,    wherein a planet carrier of the planetary gear comprises a fastening    region for attachment to a transmission housing, wherein a hollow    shaft of the planetary gear is fixedly connected to the drive shaft    is, and wherein a sun gear of the planetary gear is configured as a    hollow shaft which is arranged concentrically to the drive shaft and    rotatably mounted on the drive shaft.-   81. Cycloidal gear, the cycloidal gear comprising the following    components:    -   a transmission housing,    -   an outer gear having an internal toothing which is fixed to the        transmission housing,    -   an input shaft arranged concentric with the outer gear, wherein        the input shaft comprises a drive side eccentric disk, on which        a first ball bearing is arranged, and an output side eccentric        disk, on which a second ball bearing is arranged,    -   a drive-side inner gear with an external toothing which is        mounted on the first ball bearing and an output-side inner gear        with an external toothing which is mounted on the second ball        bearing,    -   wherein the drive side inner gear and the output side inner gear        are disposed inside the outer gear, and wherein the external        toothings of the drive side inner gear and the output-side outer        gear respectively engage with the internal toothing of the outer        gear.-   82. Cycloidal gear according to item 81, wherein the cycloidal gear    comprises a crankshaft which is mounted within the input shaft, and    a load cell according to item 63 which is mounted on the crankshaft    on the drive side.-   83. Cycloidal gear according to any of items 81 or 82, wherein the    input shaft is configured as a one-piece rotor transmitter element    according to any of items 73 to 77.-   84. Cycloidal gear according to any of items 81 to 83, wherein the    cycloidal gear comprises a crankshaft which is mounted within the    input shaft, and wherein the crankshaft comprises planetary gears    according to the items 78 to 80, wherein the crankshaft forms the    drive shaft of the planetary gear.-   85. Cycloidal gear according to any of items 81 to 84,    -   wherein a third ball bearing is disposed on the input shaft on        the output side of the output side eccentric disk, wherein a        driven pulley is arranged on the third ball bearing, wherein the        driven pulley comprises carrier pins which engage in openings of        the drive side inner gear and the output side inner gear,        wherein an output shaft is formed radially inside on the driven        pulley.-   86. Cycloidal gear according to item 85, wherein an output shaft is    formed radially inside on the driven pulley, wherein the third ball    bearing is arranged on an inner shoulder of the output shaft and    wherein an inner gear ball bearing is arranged diagonally opposite    to the third ball bearing on an outer shoulder of the output shaft,    wherein the inner gear ball bearing is supported on the transmission    housing.-   87. Cycloidal gear according to any of items 81 to 83,    -   wherein at least one of the inner gears comprises a first        toothing and a second external toothing and wherein the        cycloidal gear comprises a rotatable outer gear having an        internal toothing, wherein the second external toothing engages        in the internal toothing of the rotatable outer gear, and        wherein the rotatable outer gear comprises a fastening region        for mounting an output shaft.

The objects of the present description may also be described by theabove enumeration. The particular feature combinations disclosed in theenumeration may be considered as independent items that may also becombined with other features of the present description.

In particular, the features of FIGS. 21 to 40 relating to a torquemeasuring device, the features of FIGS. 45 to 59 relating to a freewheeldevice and the features of FIGS. 67 to 76 relating to a crank gear maybe combined with the harmonic pin ring transmissions shown in FIGS. 1 to10, 11 to 20 and 61 to 66 as well as with the other reduction gears ormotor gear units shown in FIGS. 77 to 93 and 124 to 126. These featurescan also be combined with transmissions or reduction gears not shownhere.

Furthermore, the bearing arrangement shown in FIG. 1 can also becombined with the other transmission units shown in the presentdescription. The same applies to the one-piece rotor-transmitter elementand the support ring arrangement for the outer gear or the outer gears,as well as for the one-piece pin ring, insofar as the transmissioncomprises a pin ring.

Further, example aspects of the present disclosure can comprise:

-   1. Load cell (47) for determining a radial force acting on a    crankshaft (35), the load cell (47) comprising:    -   a receiving sleeve (96) for receiving a ring of a bearing (45),    -   a fastening ring (97) for attaching the load cell (47) in a        transmission housing (22),    -   axial support areas (91) provided on the fastening ring (97) for        axially supporting the ring of the bearing (45),    -   measuring regions (90) for receiving radial forces of the        receiving sleeve (96) and which connect the receiving sleeve        (96) with the fastening ring (97), wherein strain sensors (92)        are attached to at least two of the measuring regions (90),        wherein the axial support areas (91) are separated from the        measuring regions (90) by radial slots (95), and wherein the        axial support areas (91) are each separated from the receiving        sleeve (96) by a circumferential slot (104), wherein a first        radial slot (95), a circumferential slot (104) and a second        radial slot (95) together form a confining slot which separates        an axial support area (91) from the receiving sleeve (96) and        from adjacent measuring regions (90) and wherein the axial        support areas (91) project radially inwardly over an inner        surface of the receiving sleeve (96).-   2. Load cell (47) according to item 1, wherein the measuring regions    (90) comprise measuring lugs (90) formed as angle brackets.-   3. Load cell (47) according to item 2, wherein the measuring lugs    (90) comprise a radial region and an axial region adjoining the    radial region, wherein the radial region is connected to the    fastening ring (97), wherein the axial region is connected to the    receiving sleeve (96) and wherein the radial region is arranged to    the axial region at an angle of approximately 90 degrees.-   4. Load cell (47) according to item 3, wherein the axial region is    flush with a cylindrical inner surface of the receiving sleeve (96).-   5. Load cell (47) according to any of items 1 to 4, wherein at least    one of the strain sensors (92) is configured as a strain gauge (92).-   6. Load cell (47) according to any of items 1 to 5,    -   wherein a strain sensor (92) is attached to each of the        measuring regions (90).-   7. Load cell (47) according to any of items 1 to 6,    -   wherein at least two of the measuring regions (90) comprise        lowered areas for attaching the strain sensors (92).-   8. Load cell (47) according to any of items 1 to 7,    -   wherein the load cell (47) comprises four measuring regions        (90), which are arranged at intervals of 90 degrees.-   9. Load cell (47) according to any of items 1 to 8, wherein the    fastening ring (97) comprises fastening regions in which fixing    holes (98) are provided.-   10. Load cell (47) according to any of the preceding items 1 to 9,    wherein the fastening ring (97) comprises recesses, and wherein the    measuring regions (90) are arranged in the recesses.-   11. Load cell (47) according to any of the preceding items 1 to 10,    -   wherein an angular range of the measuring lugs (90) and of the        slots (95) delimiting the measuring lugs (90) corresponds        approximately to an angular range of the axial support areas        (91).-   12. Load cell (47) according to any of the preceding items 1 to 11,    wherein an angular extent of the measuring regions (90) in a    circumferential direction is less than or equal to 30 degrees.-   13. Load cell according to any of the preceding items 1 to 12,    -   wherein the load cell (47) is made integrally of metal.-   14. Measuring device for determining a force acting on a crankshaft    (35), comprising    -   a crankshaft (35) with a bearing (45) arranged on the crankshaft        (35),    -   a load cell (47) according to any of the preceding claims,        wherein the receiving sleeve (96) of the load cell (47) is        arranged on an outer ring of the bearing (45), and wherein the        axial support areas (91) of the load cell (47) are supported in        an axial direction on the outer ring of the bearing (45),    -   an evaluation electronics (48) which is connected to the strain        sensors (90) of the load cell (47).-   15. Transmission arrangement with a measuring device according to    item 14 with a transmission housing (22) and a crankshaft (35),    wherein the crankshaft (35) is mounted in the transmission housing    (22) via a first bearing (45) and a second bearing (46),    -   wherein the first bearing (45) is received in the transmission        housing (22) via the load cell (47) of the measuring device,        wherein the load cell (47) is received in the transmission        housing (22) via the fastening ring (97), wherein the receiving        sleeve (96) receives an outer ring of the first bearing (45) and        wherein the axial support areas (91) are supported on the outer        ring of the first bearing (45).-   16. Transmission arrangement according to item 15, wherein the    crankshaft (35) comprises a first step (223) and a second step (226)    and wherein an inner ring of the first bearing (45) of the measuring    device rests against the first step (223) of the crankshaft (35) and    wherein an inner ring of the second bearing (46) rests against the    second step (226) of the crankshaft (35) such that an X-arrangement    of an obliquely mounted bearing is formed.-   17. Transmission arrangement according to item 15 or item 16,    wherein the first bearing (45) of the measuring device and the    second bearing (46) of the measuring device are each configured as    single-row angular contact ball bearings.-   18. Transmission arrangement according to any of items 15 to 17,    wherein the second bearing (46) is supported by a wave spring (70)    on the second step (226) of the crankshaft (35) or on the housing.-   19. Transmission arrangement according to item 18, wherein the    second bearing (46) is supported further by a spacer disk (69) on    the second step (226) of the crankshaft (35) or on the housing.-   20. Transmission arrangement according to any of items 15 to 19,    further comprising:    -   a motor (12), a reduction gear (13) connected to the motor (12),        and a hollow output shaft (39) connected to the reduction gear        (13),    -   wherein the crankshaft (35) is configured as pedal shaft (35),        wherein the first bearing (45) and the second bearing (46) are        each configured as rolling bearings, wherein the pedal shaft        (35) is passed through the hollow output shaft (39), and wherein        a freewheel (49) is provided between the pedal shaft (35) and        the hollow output shaft (39) for decoupling the pedal shaft (35)        from the hollow output shaft (39).-   21. Electrically driven vehicle with a transmission arrangement    according to item 20, wherein the motor (12) is configured as an    electric motor (12), and wherein a battery of the electrically    driven vehicle is connected to the electric motor (12).

In the following, items of the disclosure are expressed in structuredform. These items can be made the subject of independent or dependentclaims either alone, or in arbitrary combination among each other, andin combination with other items disclosed herein.

-   1. Load cell for determining a radial force acting on a crankshaft,    the load cell comprising:    -   a receiving sleeve for receiving a ring of a bearing,    -   a fastening ring for attaching the load cell in a transmission        housing,    -   axial support areas provided on the fastening ring for axially        supporting the ring of the bearing,    -   measuring regions for receiving radial forces of the receiving        sleeve and which connect the receiving sleeve with the fastening        ring, wherein strain sensors are attached to at least two of the        measuring regions.-   2. Load cell according to item 1, wherein the axial support areas    are separated from the measuring regions by radial slots, and    wherein the axial support areas are separated from the receiving    sleeve by a circumferential slot.-   3. Load cell according to item 1 or item 2, wherein the measuring    regions comprise measuring lugs formed as angle brackets.-   4. Load cell according to item 3, wherein the measuring lugs    comprise a radial region and an axial region adjoining the radial    region,    -   wherein the radial region is connected to the fastening ring,        wherein the axial region is connected to the receiving sleeve        and wherein the radial region is arranged to the axial region at        an angle of approximately 90 degrees.-   5. Load cell according to item 4, wherein the axial region is flush    with a cylindrical inner surface of the receiving sleeve.-   6. Load cell according to any of items 1 to 5, wherein the axial    support areas project radially inwardly over an inner surface of the    receiving sleeve.-   7. Load cell according to any of items 1 to 6, wherein at least one    of the strain sensors is configured as a strain gauge.-   8. Load cell according to any of items 1 to 7, wherein a strain    sensor is attached to each of the measuring regions.-   9. Load cell according to any of items 1 to 8, wherein at least two    of the measuring regions comprise lowered areas for attaching the    strain sensors.-   10. Load cell according to any of items 1 to 9, wherein the load    cell comprises four measuring regions, which are arranged at    intervals of 90 degrees.-   11. Load cell according to any of items 1 to 10, wherein the    fastening ring comprises fastening regions in which fixing holes are    provided.-   12. Load cell according to any of the preceding items, wherein the    fastening ring comprises recesses, and wherein the measuring regions    are arranged in the recesses.-   13. Load cell according to any of the preceding items, wherein an    angular range of the measuring lugs and of the slots delimiting the    measuring lugs corresponds approximately to an angular range of the    axial support areas.-   14. Load cell according to any of the preceding items, wherein an    angular extent of the measuring regions in a circumferential    direction is less than or equal to 30 degrees.-   15. Load cell according to any of the preceding items, wherein the    load cell is made integrally of metal.-   16. Measuring device for determining a force acting on a crankshaft,    comprising    -   a crankshaft with a bearing arranged on the crankshaft,    -   a load cell according to any of the preceding items, wherein the        receiving sleeve of the load cell is arranged on an outer ring        of the bearing, and wherein the axial support areas of the load        cell are supported in an axial direction on the outer ring of        the bearing,    -   an evaluation electronics which is connected to the strain        sensors of the load cell.-   17. Transmission arrangement with a measuring device according to    item 16 with a transmission housing and a crankshaft, wherein the    crankshaft is mounted in the transmission housing via a first    bearing and a second bearing,    -   wherein the first bearing is received in the transmission        housing via the load cell of the measuring device, wherein the        load cell is received in the transmission housing via the        fastening ring, wherein the receiving sleeve receives an outer        ring of the first bearing and wherein the axial support areas        are supported on the outer ring of the first bearing.-   18. Transmission arrangement according to item 17, wherein the    crankshaft comprises a first step and a second step and wherein an    inner ring of the first bearing of the measuring device rests    against the first step of the pedal shaft and wherein an inner ring    of the second bearing rests against the second step of the    crankshaft such that an X-arrangement of an obliquely mounted    bearing is formed.-   19. Transmission arrangement according to item 17 or item 18,    wherein the first bearing of the measuring device and the second    bearing of the measuring device are each configured as single-row    angular contact ball bearings.-   20. Transmission arrangement according to any of items 17 to 19,    wherein the second bearing is supported by a wave spring on the    second step of the pedal shaft or on the housing.-   21. Transmission arrangement according to item 20, wherein the    second bearing is supported further by a spacer disk on the second    step of the pedal shaft or on the housing.-   22. Transmission arrangement according to any of items 17 to 21,    further comprising:    -   a motor, a reduction gear connected to the motor, and a hollow        output shaft connected to the reduction gear,    -   wherein the crankshaft is configured as pedal shaft, wherein the        first bearing and the second bearing are each configured as        rolling bearings, wherein the pedal shaft is passed through the        hollow output shaft, and wherein a freewheel is provided between        the pedal shaft and the hollow output shaft for decoupling the        pedal shaft from the hollow output shaft.-   23. Electrically driven vehicle with a transmission arrangement    according to item 22, wherein the motor is configured as an electric    motor, and wherein a battery of the electrically driven vehicle is    connected to the electric motor.

Reference numerals   5 external toothing   5′ external toothing   6internal toothing   6′ internal toothing   7 inner gear   8 outer gear  8′ outer gear   9 step  10 motor gear unit  11 step  12 motor  13reduction gear  20 stator  21 coil  22 motor housing  23 printed circuitboard  24 cooling cover  25 terminal  26 outer rotor shaft  27 innerrotor shaft  28 cam disk  28′ eccentric disk  29 drive side rotor ballbearing  30 output side rotor ball bearing  31 inner gear ball bearing 32 housing cover  33 flexible ball bearing  33′ ball bearing  34fastening screws  35 pedal shaft  36 support ring  37 drive side spacerdisk  38 output side spacer disk  39 output shaft  40 transmissionfreewheel  41 output ball bearing  41′ drive side output ball bearing 42 O-ring  43 chainring adapter  44 output nut, transmission cover  45drive side pedal shaft ball bearing  46 output side pedal shaft ballbearing  47 load cell  48 annular printed circuit board/PCB   forcesensor  49 pedal shaft freewheel  50 inner shaft seal ring  51 outershaft seal ring  52 drive side shaft seal ring  53 Journal  54 radialslots  55 screw holes  56 transmitting region  57 step  58 roundrecesses  59 plateaus  60 screw thread  61 wave spring  62 spacer ring 63 ribbon cable  64 cylindrical rollers  65 clamping body cage  66 coilsprings  67 raceway  68 sensor ring  69 spacer  70 wave spring  71 steps 72 coil springs  73 pawls  74 fastening region  75 fastening region  76O-ring  80 pedal shaft assembly  81 freewheel assembly  82 sensorassembly  90 measuring lug  91 support lug  92 strain gauge  93 outerportion  94 inner portion  95 radial slot  96 sleeve  97 outer ring  98fastening holes  99 recessed area 101 pins 102 pin ring 103 pinretaining ring 104 slots 105 circumferential slots 106 drive side step107 output side step 109 bottom bracket bearing 110 torque measuringdevice 111 pedal shaft 112 first rolling bearing 113 second rollingbearing 114 sleeve 115 end face 116 portion 117 strain gauges 118 straingauges 119 strain gauges 120 outer surface 121 slip rings 122transmitter 124 strain gauges 130 torque measuring device 131 straingauges 132 strain gauges 133 strain gauges 134 strain gauges 135 housing136 printed circuit board 137 transmitter ring 138 sensor 139 protrusion140 thread 141 step 142 O-ring 220 receiving region 222 step, drive side223 step, drive side 224 step, output side 225 step, output side 226step, output side 227 star arrangement 228 rolling region 229 rollingregion 230 step 231 step 232 step 233 internal thread 234 hollow shaftportion 235 disk-shaped region 236 bore 237 chamfer 243 rolling region244 end portion 245 pawls receiving region 246 hinge portion 247 hingeportion 248 plate-shaped portion 249 edge 250 webs 251 receiving region252 receiving region 255 first region 257 sensor ring 258 rotation speedsensor 262 O-ring 343 O-ring 344 O-ring 345 sensor ring 346 innerhousing 347 screws 348 support cylinder 349 webs 352 hollow shaft 353Hall sensor 354 grooves 355 journals 400, 400′ planetary gear assembly401, 401′ ring gear 402 fastening flange 403 sun gear 405 planet gear406 planetary axis 407 planet carrier 408 rolling bearing 409 hollowshaft 410 hollow shaft 423 ball bearing 424 ball bearing 425 ballbearing 426 drive shaft 427 rotor shaft 428 eccentric disk 429 eccentricdisk 430 centered circular disk 433 inner gear 434 inner gear 435external toothing 436 external toothing 437 internal toothing 440 outputpulley 442 carrier pins 443 carrier rollers 441 ring 444 circularopening 445 spacer 446 ridge 447 stationary pins 448 fastening opening450 external thread 451 pressure disk 452 pressure ring 453 tensioningshaft 454 fastening region 455 drive cylinder 456 fastening flange 457fastening region 458 output shaft 501 pins 502 flexible thin sectionball bearing 503 pin retaining ring 504 flange 505, 505′ inner geartoothing 506, 506′ outer gear toothing 507 Inner ring or inner gear 508outer ring or outer gear 509 cylindrical housing part 510HPRD-transmission 513 rotor 514 eccentric disk 515 transmission centralaxis 516 semicircular recesses 518 gear trajectory 519 equidistant 520pin cross section 521 position of the gear trajectory 525 zero contour526 contour with provision 527 worn out contour 528 gear trajectory ofthe inner gear 528′ gear trajectory of the outer gear 529 equidistant ofthe gear trajectory of the inner gear 529′ equidistant of the geartrajectory of the outer gear 530 integrally formed pin ring 531 tooth ofthe pin ring 532 tooth base of the pin ring 533 inner envelope curve 534outer envelope curve 536 predetermined tooth profile 537 tooth profileto be measured

That which is claimed is:
 1. A measuring device for determining a forceacting on a crankshaft, the measuring device comprising: the crankshaftwith a bearing arranged on the crankshaft; and a load cell fordetermining a radial force acting on the crankshaft, the load cellcomprising: a receiving sleeve for receiving a ring of the bearing; afastening ring for attaching the load cell in a transmission housing;axial support areas provided on the fastening ring for axiallysupporting the ring of the bearing; and measuring regions for receivingradial forces of the receiving sleeve and which connect the receivingsleeve with the fastening ring, wherein strain sensors are attached toat least two of the measuring regions; wherein the receiving sleeve ofthe load cell is arranged on an outer ring of the bearing, and whereinthe axial support areas of the load cell are supported in an axialdirection on the outer ring of the bearing; and wherein an evaluationelectronics is connected to the strain sensors of the load cell.
 2. Themeasuring device of claim 1, wherein the axial support areas areseparated from the measuring regions by radial slots, and wherein theaxial support areas are separated from the receiving sleeve by acircumferential slot.
 3. The measuring device of claim 1, wherein themeasuring regions comprise measuring lugs formed as angle brackets. 4.The measuring device of claim 3, wherein: the measuring lugs comprise aradial region and an axial region adjoining the radial region; theradial region is connected to the fastening ring; the axial region isconnected to the receiving sleeve; and the radial region is arranged tothe axial region at an angle of approximately 90 degrees.
 5. Themeasuring device of claim 4, wherein the axial region is flush with acylindrical inner surface of the receiving sleeve.
 6. The measuringdevice of claim 1, wherein the axial support areas project radiallyinwardly over an inner surface of the receiving sleeve.
 7. The measuringdevice of claim 1, wherein at least one of the strain sensors isconfigured as a strain gauge.
 8. The measuring device of claim 1,wherein a strain sensor is attached to each of the measuring regions. 9.The measuring device of claim 1, wherein at least two of the measuringregions comprise lowered areas for attaching the strain sensors.
 10. Themeasuring device of claim 1, wherein the load cell comprises fourmeasuring regions, which are arranged at intervals of 90 degrees. 11.The measuring device of claim 1, wherein the fastening ring comprisesfastening regions in which fixing holes are provided.
 12. The measuringdevice of claim 1, wherein the fastening ring comprises recesses, andwherein the measuring regions are arranged in the recesses.
 13. Themeasuring device of claim 1, wherein an angular range of the measuringlugs and of slots delimiting the measuring lugs correspondsapproximately to an angular range of the axial support areas.
 14. Themeasuring device of claim 1, wherein an angular extent of the measuringregions in a circumferential direction is less than or equal to 30degrees.
 15. The measuring device of claim 1, wherein the load cell ismade integrally of metal.
 16. A transmission arrangement comprising: ameasuring device comprising: a crankshaft with a bearing arranged on thecrankshaft; and a load cell for determining a radial force acting on thecrankshaft, the load cell comprising: a receiving sleeve for receiving aring of the bearing; a fastening ring for attaching the load cell in atransmission housing; axial support areas provided on the fastening ringfor axially supporting the ring of the bearing; and measuring regionsfor receiving radial forces of the receiving sleeve and which connectthe receiving sleeve with the fastening ring, wherein strain sensors areattached to at least two of the measuring regions; wherein the receivingsleeve of the load cell is arranged on an outer ring of the bearing, andwherein the axial support areas of the load cell are supported in anaxial direction on the outer ring of the bearing; and wherein anevaluation electronics is connected to the strain sensors of the loadcell; wherein the crankshaft is mounted in the transmission housing viaa first bearing and a second bearing, wherein the first bearing isreceived in the transmission housing via the load cell of the measuringdevice; wherein the load cell is received in the transmission housingvia the fastening ring; wherein the receiving sleeve receives an outerring of the first bearing; and wherein the axial support areas aresupported on the outer ring of the first bearing.
 17. The transmissionarrangement of claim 16, wherein the crankshaft comprises a first stepand a second step and wherein an inner ring of the first bearing of themeasuring device rests against the first step of the crankshaft andwherein an inner ring of the second bearing rests against the secondstep of the crankshaft such that an X-arrangement of an obliquelymounted bearing is formed.
 18. The transmission arrangement of claim 17,further comprising a motor, a reduction gear connected to the motor, anda hollow output shaft connected to the reduction gear, and wherein: thecrankshaft is configured as a pedal shaft; the first bearing and thesecond bearing are each configured as a rolling bearing; the pedal shaftis passed through the hollow output shaft; and a freewheel is providedbetween the pedal shaft and the hollow output shaft for decoupling thepedal shaft from the hollow output shaft.
 19. The transmissionarrangement according to claim 18, wherein: the transmission arrangementis configured for use with an electrically driven vehicle; the motor isconfigured as an electric motor; and a battery of the electricallydriven vehicle is connected to the electric motor.
 20. The transmissionarrangement of claim 16, wherein the first bearing of the measuringdevice and the second bearing of the measuring device are eachconfigured as single-row angular contact ball bearings.
 21. Thetransmission arrangement of claim 16, wherein the second bearing issupported by a wave spring on the second step of the crankshaft or onthe transmission housing.
 22. The transmission arrangement of claim 21,wherein the second bearing is supported further by a spacer disk on thesecond step of the crankshaft or on the transmission housing.